Compressor with liquid injection cooling

ABSTRACT

A compressor includes: a casing with an inner wall defining a compression chamber, an inlet leading into the compression chamber, and an outlet leading out of the compression chamber; a rotor rotatably coupled to the casing for rotation relative to the casing; and a gate coupled to the casing for movement relative to the casing. The gate may be pivotally, or translationally coupled to the casing. A hydrostatic bearing may be disposed between the gate and casing. A plurality of compressors may be mechanically linked together such that their compression cycles are out of phase.

1. CROSS-REFERENCE TO RELATED APPLICATION

This application claims the benefit of U.S. Provisional Application Ser.No. 62/139,884, filed on Mar. 30, 2015, the content of which is herebyincorporated herein by reference in its entirety.

BACKGROUND 1. Technical Field

The invention generally relates to fluid pumps, such as compressors andexpanders.

2. Related Art

Compressors have typically been used for a variety of applications, suchas air compression, vapor compression for refrigeration, and compressionof industrial gases. Compressors can be split into two main groups,positive displacement and dynamic. Positive displacement compressorsreduce the compression volume in the compression chamber to increase thepressure of the fluid in the chamber. This is done by applying force toa drive shaft that is driving the compression process. Dynamiccompressors work by transferring energy from a moving set of blades tothe working fluid.

Positive displacement compressors can take a variety of forms. They aretypically classified as reciprocating or rotary compressors.Reciprocating compressors are commonly used in industrial applicationswhere higher pressure ratios are necessary. They can easily be combinedinto multistage machines, although single stage reciprocatingcompressors are not typically used at pressures above 80 psig.Reciprocating compressors use a piston to compress the vapor, air, orgas, and have a large number of components to help translate therotation of the drive shaft into the reciprocating motion used forcompression. This can lead to increased cost and reduced reliability.Reciprocating compressors also suffer from high levels of vibration andnoise. This technology has been used for many industrial applicationssuch as natural gas compression.

Rotary compressors use a rotating component to perform compression. Asnoted in the art, rotary compressors typically have the followingfeatures in common: (1) they impart energy to the gas being compressedby way of an input shaft moving a single or multiple rotating elements;(2) they perform the compression in an intermittent mode; and (3) theydo not use inlet or discharge valves. (Brown, Compressors: Selection andSizing, 3rd Ed., at 6). As further noted in Brown, rotary compressordesigns are generally suitable for designs in which less than 20:1pressure ratios and 1000 CFM flow rates are desired. For pressure ratiosabove 20:1, Royce suggests that multistage reciprocating compressorsshould be used instead.

Typical rotary compressor designs include the rolling piston, screwcompressor, scroll compressor, lobe, liquid ring, and rotary vanecompressors. Each of these traditional compressors has deficiencies forproducing high pressure, near isothermal conditions.

The design of a rotating element/rotor/lobe against a radially movingelement/piston to progressively reduce the volume of a fluid has beenutilized as early as the mid-19th century with the introduction of the“Yule Rotary Steam Engine.” Developments have been made to small-sizedcompressors utilizing this methodology into refrigeration compressionapplications. However, current Yule-type designs are limited due toproblems with mechanical spring durability (returning the pistonelement) as well as chatter (insufficient acceleration of the piston inorder to maintain contact with the rotor).

For commercial applications, such as compressors for refrigerators,small rolling piston or rotary vane designs are typically used. (P NAnanthanarayanan, Basic Refrigeration and Air Conditioning, 3rd Ed., at171-72.) In these designs, a closed oil-lubricating system is typicallyused.

Rolling piston designs typically allow for a significant amount ofleakage between an eccentrically mounted circular rotor, the interiorwall of the casing, and/or the vane that contacts the rotor. By spinningthe rolling piston faster, the leakages are deemed acceptable becausethe desired pressure and flow rate for the application can be easilyreached even with these losses. The benefit of a small self-containedcompressor is more important than seeking higher pressure ratios.

Rotary vane designs typically use a single circular rotor mountedeccentrically in a cylinder slightly larger than the rotor. Multiplevanes are positioned in slots in the rotor and are kept in contact withthe cylinder as the rotor turns typically by spring or centrifugal forceinside the rotor. The design and operation of these type of compressorsmay be found in Mark's Standard Handbook for Mechanical Engineers,Eleventh Edition, at 14:33-34.

In a sliding-vane compressor design, vanes are mounted inside the rotorto slide against the casing wall. Alternatively, rolling piston designsutilize a vane mounted within the cylinder that slides against therotor. These designs are limited by the amount of restoring force thatcan be provided and thus the pressure that can be yielded.

Each of these types of prior art compressors has limits on the maximumpressure differential that it can provide. Typical factors includemechanical stresses and temperature rise. One proposed solution is touse multistaging. In multistaging, multiple compression stages areapplied sequentially. Intercooling, or cooling between stages, is usedto cool the working fluid down to an acceptable level to be input intothe next stage of compression. This is typically done by passing theworking fluid through a heat exchanger in thermal communication with acooler fluid. However, intercooling can result in some condensation ofliquid and typically requires filtering out of the liquid elements.Multistaging greatly increases the complexity of the overall compressionsystem and adds costs due to the increased number of componentsrequired. Additionally, the increased number of components leads todecreased reliability and the overall size and weight of the system aremarkedly increased.

For industrial applications, single- and double-acting reciprocatingcompressors and helical-screw type rotary compressors are most commonlyused. Single-acting reciprocating compressors are similar to anautomotive type piston with compression occurring on the top side of thepiston during each revolution of the crankshaft. These machines canoperate with a single-stage discharging between 25 and 125 psig or intwo stages, with outputs ranging from 125 to 175 psig or higher.Single-acting reciprocating compressors are rarely seen in sizes above25 HP. These types of compressors are typically affected by vibrationand mechanical stress and require frequent maintenance. They also sufferfrom low efficiency due to insufficient cooling.

Double-acting reciprocating compressors use both sides of the piston forcompression, effectively doubling the machine's capacity for a givencylinder size. They can operate as a single-stage or with multiplestages and are typically sized greater than 10 HP with dischargepressures above 50 psig. Machines of this type with only one or twocylinders require large foundations due to the unbalanced reciprocatingforces. Double-acting reciprocating compressors tend to be quite robustand reliable, but are not sufficiently efficient, require frequent valvemaintenance, and have extremely high capital costs.

Lubricant-flooded rotary screw compressors operate by forcing fluidbetween two intermeshing rotors within a housing which has an inlet portat one end and a discharge port at the other. Lubricant is injected intothe chamber to lubricate the rotors and bearings, take away the heat ofcompression, and help to seal the clearances between the two rotors andbetween the rotors and housing. This style of compressor is reliablewith few moving parts. However, it becomes quite inefficient at higherdischarge pressures (above approximately 200 psig) due to theintermeshing rotor geometry being forced apart and leakage occurring. Inaddition, lack of valves and a built-in pressure ratio leads to frequentover or under compression, which translates into significant energyefficiency losses.

Rotary screw compressors are also available without lubricant in thecompression chamber, although these types of machines are quiteinefficient due to the lack of lubricant helping to seal between therotors. They are a requirement in some process industries such as foodand beverage, semiconductor, and pharmaceuticals, which cannot tolerateany oil in the compressed air used in their processes. Efficiency of dryrotary screw compressors are 15-20% below comparable injected lubricatedrotary screw compressors and are typically used for discharge pressuresbelow 150 psig.

Using cooling in a compressor is understood to improve upon theefficiency of the compression process by extracting heat, allowing mostof the energy to be transmitted to the gas and compressing with minimaltemperature increase. Liquid injection has previously been utilized inother compression applications for cooling purposes. Further, it hasbeen suggested that smaller droplet sizes of the injected liquid mayprovide additional benefits.

In U.S. Pat. No. 4,497,185, lubricating oil was intercooled and injectedthrough an atomizing nozzle into the inlet of a rotary screw compressor.In a similar fashion, U.S. Pat. No. 3,795,117 uses refrigerant, thoughnot in an atomized fashion, that is injected early in the compressionstages of a rotary screw compressor. Rotary vane compressors have alsoattempted finely atomized liquid injection, as seen in U.S. Pat. No.3,820,923.

Published International Pat. App. No. WO 2010/017199 and U.S. Pat. Pub.No. 2011/0023814 relate to a rotary engine design using a rotor,multiple gates to create the chambers necessary for a combustion cycle,and an external cam-drive for the gates. The force from the combustioncycle drives the rotor, which imparts force to an external element.Engines are designed for a temperature increase in the chamber and hightemperatures associated with the combustion that occurs within anengine. Increased sealing requirements necessary for an effectivecompressor design are unnecessary and difficult to achieve. Combustionforces the use of positively contacting seals to achieve near perfectsealing, while leaving wide tolerances for metal expansion, taken up bythe seals, in an engine. Further, injection of liquids for cooling wouldbe counterproductive and coalescence is not addressed.

Liquid mist injection has been used in compressors, but with limitedeffectiveness. In U.S. Pat. No. 5,024,588, a liquid injection mist isdescribed, but improved heat transfer is not addressed. In U.S. Pat.Publication. No. U.S. 2011/0023977, liquid is pumped through atomizingnozzles into a reciprocating piston compressor's compression chamberprior to the start of compression. It is specified that liquid will onlybe injected through atomizing nozzles in low pressure applications.Liquid present in a reciprocating piston compressor's cylinder causes ahigh risk for catastrophic failure due to hydrolock, a consequence ofthe incompressibility of liquids when they build up in clearance volumesin a reciprocating piston, or other positive displacement, compressor.To prevent hydrolock situations, reciprocating piston compressors usingliquid injection will typically have to operate at very slow speeds,adversely affecting the performance of the compressor.

U.S. Patent Application Publication No. 2013-0209299, titled “CompressorWith Liquid Injection Cooling” discloses another rotary compressor withliquid injection cooling. The entire contents of U.S. Patent ApplicationPublication No. 2013-0209299 are incorporated herein by reference in itsentirety.

BRIEF SUMMARY

The presently preferred embodiments are directed to rotary compressordesigns. These designs are particularly suited for high pressureapplications, typically above 200 psig with pressure ratios typicallyabove that for existing high-pressure positive displacement compressors.

One or more embodiments provides a compressor that includes: a casingwith an inner wall defining a compression chamber; a drive shaft androtor rotatably coupled to the casing for common rotation relative tothe casing, the rotor having a non-circular profile; and a gate coupledto the casing for pivotal movement relative to the casing, the gatecomprising a sealing edge, the gate being operable to move relative tothe casing to locate the sealing edge proximate to the rotor as therotor rotates such that the gate separates an inlet volume and acompression volume in the compression chamber.

One or more embodiments provides a compressor that includes: a casingwith an inner wall defining a compression chamber, an inlet leading intothe compression chamber, and an outlet leading out of the compressionchamber; a drive shaft and rotor rotatably coupled to the casing forcommon rotation relative to the casing, the rotor having a non-circularprofile; a gate coupled to the casing for movement relative to thecasing, the gate comprising a sealing edge, the gate being operable tomove relative to the casing to locate the sealing edge proximate to therotor as the rotor rotates such that the gate separates an inlet volumeand a compression volume in the compression chamber, the inlet andoutlet being disposed on opposite sides of the sealing edge from eachother; and an outlet manifold in fluid communication with the outlet,wherein the outlet is elongated in a direction parallel to a rotationalaxis of the drive shaft, wherein the outlet manifold defines an interiorpassageway, and wherein the passageway varies in cross-sectional shapebetween an entrance into the manifold and an exit out of the manifold,and wherein the outlet manifold comprises a plurality of vanes disposedin the interior passageway to direct the flow of working fluid throughthe outlet manifold.

One or more embodiments provides a compressor that includes: a casingwith an inner wall defining a compression chamber, an inlet leading intothe compression chamber, and an outlet leading out of the compressionchamber; a rotor coupled to the casing for rotation relative to thecasing; a gate movably coupled to one of the casing and rotor formovement relative to the one of the casing and rotor, the gatecomprising a sealing edge, the gate being operable to locate the sealingedge proximate to the other of the casing and rotor as the rotorrotates; and a hydrostatic bearing arrangement disposed between (1) thegate and (2) the one of the casing and rotor to reduce friction when thegate moves during operation of the compressor.

One or more embodiments provides a compressor that includes: acompression chamber casing with an inner wall defining a compressionchamber, an inlet leading into the compression chamber, and an outletleading out of the compression chamber; a drive shaft and rotorrotatably coupled to the compression chamber casing for common rotationrelative to the compression chamber casing; a gate coupled to thecompression chamber casing for movement relative to the compressionchamber casing, the gate comprising a sealing edge, the gate beingoperable to move relative to the compression chamber casing to locatethe sealing edge proximate to the rotor as the rotor rotates such thatthe gate separates an inlet volume and a compression volume in thecompression chamber, the inlet and outlet being disposed on oppositesides of the sealing edge from each other; and a gate positioning systemcoupled to the gate, the gate positioning system being shaped andconfigured to reciprocally move the gate during rotation of the rotor sothat the sealing edge remains proximate to the rotor during rotation ofthe rotor.

According to various embodiments, the gate positioning system includes acam shaft rotatably coupled to the compression chamber casing forrotation relative to the compression chamber casing, the cam shaft beingspaced from the drive shaft, the cam shaft being connected to the driveshaft so as to be rotationally driven by the drive shaft, a camrotatably coupled to the compression chamber casing for concentricrotation with the cam shaft relative to the compression chamber casing,a cam follower mounted to the gate for movement with the gate relativeto the compression chamber casing, the cam follower abutting the cam sothat rotation of the cam causes the cam follower and gate to moverelative to the compression chamber casing.

One or more embodiments provides a compressor system that includes: aplurality of compressors. Each compressor may include a casing with aninner wall defining a compression chamber, an inlet leading into thecompression chamber, and an outlet leading out of the compressionchamber, a rotor rotatably coupled to the casing for rotation relativeto the casing, and a gate coupled to the casing for movement relative tothe casing, the gate comprising a sealing edge, the gate being operableto move relative to the casing to locate the sealing edge proximate tothe rotor as the rotor rotates such that the gate separates an inletvolume and a compression volume in the compression chamber, the inletand outlet being disposed on opposite sides of the sealing edge fromeach other. The system includes a mechanical linkage between the rotorsof the plurality of compressors, the mechanical linkage connectingbetween the rotors such that compression cycles of the plurality ofcompressors are out of phase with each other.

One or more embodiments provides a compressor that includes: a casingwith an inner wall defining a compression chamber, an inlet leading intothe compression chamber, and an outlet leading out of the compressionchamber; a drive shaft and rotor rotatably coupled to the casing forcommon rotation relative to the casing such that when the rotor isrotated, the compressor compresses working fluid that enters thecompression chamber from the inlet, and forces compressed working fluidout of the compression chamber through the outlet; and a mechanical seallocated at an interface between the drive shaft and casing where thedrive shaft passes through the casing.

According to various embodiments, the mechanical seal includes: first,second, and third seals disposed sequentially along a leakage pathbetween the drive shaft and casing rotor, a source of pressurizedhydraulic fluid, and a hydraulic fluid passageway that connects thesource to a space along the leakage path between the second and thirdseals so as to keep the space pressurized with hydraulic fluid.

One or more embodiments provides a non-circular seal for sealing aninterface between two moving parts. The seal includes a non-circularstructural base (e.g., comprising steel) having a closed perimeter; anda low friction sealing material (e.g., graphite or Teflon) bonded to thebase.

One or more embodiments provides a compressor that includes: a casingwith an inner wall defining a compression chamber, an inlet leading intothe compression chamber, and an outlet leading out of the compressionchamber; a rotor rotatably coupled to the casing for rotation relativeto the casing such that when the rotor is rotated, the compressorcompresses working fluid that enters the compression chamber from theinlet, and forces compressed working fluid out of the compressionchamber through the outlet; a gate coupled to the casing forreciprocating movement relative to the casing, the gate comprising asealing edge, the gate being operable to move relative to the casing tolocate the sealing edge proximate to the rotor as the rotor rotates suchthat the gate separates an inlet volume and a compression volume in thecompression chamber; and a mechanical seal located at an interfacebetween the gate and casing. The mechanical seal includes: first,second, and third seals disposed sequentially along a leakage pathbetween the gate and casing, a source of pressurized hydraulic fluid,and a hydraulic fluid passageway that connects the source to a spacealong the leakage path between the second and third seals so as to keepthe space pressurized with hydraulic fluid.

According to various embodiments, the mechanical seal further includes avent disposed between the first and second seals, the vent being fluidlyconnected to the inlet so as to direct working fluid that leaks from thecompression chamber past the first seal back to the inlet.

According to various embodiments, the first, second, and third seals areall supported by a removable housing, such that the first, second, andthird seals and housing can be installed into the casing as a singleunit.

According to various embodiments, the mechanical seal comprises nsequential seals along the leakage path between the gate and casing,wherein 3≤n≤50, wherein n includes the first, second, and third seals,wherein one or more spaces between adjacent ones of the seals are filledwith pressurized hydraulic fluid, and wherein one or more spaces betweenadjacent ones of the seals comprise a vent that is fluidly connected onthe inlet.

These and other aspects of various non-limiting embodiments of thepresent invention, as well as the methods of operation and functions ofthe related elements of structure and the combination of parts andeconomies of manufacture, will become more apparent upon considerationof the following description and the appended claims with reference tothe accompanying drawings, all of which form a part of thisspecification, wherein like reference numerals designate correspondingparts in the various figures. In one embodiment of the invention, thestructural components illustrated herein are drawn to scale. It is to beexpressly understood, however, that the drawings are for the purpose ofillustration and description only and are not intended as a definitionof the limits of the invention. In addition, it should be appreciatedthat structural features shown or described in any one embodiment hereincan be used in other embodiments as well. As used in the specificationand in the claims, the singular form of “a”, “an”, and “the” includeplural referents unless the context clearly dictates otherwise.

All closed-ended (e.g., between A and B) and open-ended (greater than C)ranges of values disclosed herein explicitly include all ranges thatfall within or nest within such ranges. For example, a disclosed rangeof 1-10 is understood as also disclosing, among other ranged, 2-10, 1-9,3-9, etc.

BRIEF DESCRIPTION OF THE DRAWINGS

Embodiments of the invention can be better understood with reference tothe following drawings and description. The components in the figuresare not necessarily to scale, emphasis instead being placed uponillustrating the principles of various embodiments of the invention.Moreover, in the figures, like referenced numerals designatecorresponding parts throughout the different views.

FIG. 1 is a perspective view of a rotary compressor with a spring-backedcam drive in accordance with an embodiment of the present invention.

FIG. 2 is a right-side view of a rotary compressor with a spring-backedcam drive in accordance with an embodiment of the present invention.

FIG. 3 is a left-side view of a rotary compressor with a spring-backedcam drive in accordance with an embodiment of the present invention.

FIG. 4 is a front view of a rotary compressor with a spring-backed camdrive in accordance with an embodiment of the present invention.

FIG. 5 is a back view of a rotary compressor with a spring-backed camdrive in accordance with an embodiment of the present invention.

FIG. 6 is a top view of a rotary compressor with a spring-backed camdrive in accordance with an embodiment of the present invention.

FIG. 7 is a bottom view of a rotary compressor with a spring-backed camdrive in accordance with an embodiment of the present invention.

FIG. 8 is a cross-sectional view of a rotary compressor with aspring-backed cam drive in accordance with an embodiment of the presentinvention.

FIG. 9 is a perspective view of rotary compressor with a belt-driven,spring-biased gate positioning system in accordance with an embodimentof the present invention.

FIG. 10 is a perspective view of a rotary compressor with a dual camfollower gate positioning system in accordance with an embodiment of thepresent invention.

FIG. 11 is a right-side view of a rotary compressor with a dual camfollower gate positioning system in accordance with an embodiment of thepresent invention.

FIG. 12 is a left-side view of a rotary compressor with a dual camfollower gate positioning system in accordance with an embodiment of thepresent invention.

FIG. 13 is a front view of a rotary compressor with a dual cam followergate positioning system in accordance with an embodiment of the presentinvention.

FIG. 14 is a back view of a rotary compressor with a dual cam followergate positioning system in accordance with an embodiment of the presentinvention.

FIG. 15 is a top view of a rotary compressor with a dual cam followergate positioning system in accordance with an embodiment of the presentinvention.

FIG. 16 is a bottom view of a rotary compressor with a dual cam followergate positioning system in accordance with an embodiment of the presentinvention.

FIG. 17 is a cross-sectional view of a rotary compressor with a dual camfollower gate positioning system in accordance with an embodiment of thepresent invention.

FIG. 18 is perspective view of a rotary compressor with a belt-drivengate positioning system in accordance with an embodiment of the presentinvention.

FIG. 19 is perspective view of a rotary compressor with an offset gateguide positioning system in accordance with an embodiment of the presentinvention.

FIG. 20 is a right-side view of a rotary compressor with an offset gateguide positioning system in accordance with an embodiment of the presentinvention.

FIG. 21 is a front view of a rotary compressor with an offset gate guidepositioning system in accordance with an embodiment of the presentinvention.

FIG. 22 is a cross-sectional view of a rotary compressor with an offsetgate guide positioning system in accordance with an embodiment of thepresent invention.

FIG. 23 is perspective view of a rotary compressor with a linearactuator gate positioning system in accordance with an embodiment of thepresent invention.

FIGS. 24A and B are right side and cross-section views, respectively, ofa rotary compressor with a magnetic drive gate positioning system inaccordance with an embodiment of the present invention

FIG. 25 is perspective view of a rotary compressor with a scotch yokegate positioning system in accordance with an embodiment of the presentinvention.

FIGS. 26A-F are cross-sectional views of the inside of an embodiment ofa rotary compressor with a contacting tip seal in a compression cycle inaccordance with an embodiment of the present invention.

FIGS. 27A-F are cross-sectional views of the inside of an embodiment ofa rotary compressor without a contacting tip seal in a compression cyclein accordance with another embodiment of the present invention.

FIG. 28 is perspective, cross-sectional view of a rotary compressor inaccordance with an embodiment of the present invention.

FIG. 29 is a left-side view of an additional liquid injectors embodimentof the present invention.

FIG. 30 is a cross-section view of a rotor design in accordance with anembodiment of the present invention.

FIGS. 31A-D are cross-sectional views of rotor designs in accordancewith various embodiments of the present invention.

FIGS. 32A and B are perspective and right-side views of a drive shaft,rotor, and gate in accordance with an embodiment of the presentinvention.

FIG. 33 is a perspective view of a gate with exhaust ports in accordancewith an embodiment of the present invention.

FIGS. 34A and B are a perspective view and magnified view of a gate withnotches, respectively, in accordance with an embodiment of the presentinvention.

FIG. 35 is a cross-sectional, perspective view a gate with a rolling tipin accordance with an embodiment of the present invention.

FIG. 36 is a cross-sectional front view of a gate with a liquidinjection channel in accordance with an embodiment of the presentinvention.

FIG. 37 is a graph of the pressure-volume curve achieved by a compressoraccording to one or more embodiments of the present invention relativeto adiabatic and isothermal compression.

FIGS. 38(a)-(d) show the sequential compression cycle and liquid coolantinjection locations, directions, and timing according to one or moreembodiments of the invention.

FIG. 39 is a perspective view of a compressor according to analternative embodiment.

FIG. 40 is a cross-sectional view of the compressor in FIG. 39, takenalong an axis of the compressor's drive shaft.

FIG. 41 is an exploded view of the compressor in FIG. 39.

FIG. 42 is an end view of the compressor in FIG. 39.

FIG. 43 is a cross-sectional view of the compressor in FIG. 39, taken ina plane that is perpendicular to a drive shaft of the compressor

FIG. 44 is a perspective view of the view in FIG. 43 of the compressorin FIG. 39.

FIG. 45 is cross-sectional view of a discharge manifold of thecompressor in FIG. 39.

FIG. 46 is perspective view of the discharge manifold in FIG. 45.

FIG. 47 is an end view of the discharge manifold in FIG. 45.

FIG. 48 is partial, cross-sectional, perspective view of the compressorin FIG. 39, showing the hydrostatic bearing arrangement.

FIG. 49 is perspective view of the hydrostatic bearings and gate of thecompressor in FIG. 39.

FIG. 50 is diagrammatic view of the hydrostatic bearing arrangement ofthe compressor in FIG. 39.

FIG. 51 is a resistance flow diagram of the hydrostatic bearings of thecompressor in FIG. 39.

FIG. 52 is a partial cross-sectional view of FIG. 40.

FIG. 53 is a partial cross-sectional view of a compressor according toan alternative embodiment.

FIG. 54 is an enlarged, partial, cross-sectional view of FIG. 52.

FIG. 55 is a perspective view of a compressor according to analternative embodiment, with a cam casing removed to display internalcomponents.

FIG. 56 is a cross-sectional view of the compressor in FIG. 55, taken ina plane that is perpendicular to a drive shaft of the compressor.

FIG. 57 is a cross-sectional view of the compressor in FIG. 55, takenalong an axis of the compressor's drive shaft.

FIG. 58 is a perspective view of the compressor in FIG. 55, showing acam casing.

FIG. 59 is a perspective view of a compressor according to analternative embodiment.

FIG. 60 is a cross-sectional view of the compressor in FIG. 59, takenalong an axis of the compressor's drive shaft.

FIGS. 61 and 62 are cross-sectional views of a compressor according toan alternative embodiment, with the cross-sections taken perpendicularto an axis of a drive shaft of the compressor.

FIGS. 63-65 are end views of the compressor of FIGS. 61 and 62, taken atdifferent points in the compression cycle.

FIG. 66 is a cross-sectional view of a compressor according to analternative embodiment, taken along an axis of the compressor's driveshaft.

FIG. 67 is a cross-sectional end view of the rotor of the compressor inFIG. 39, with the cross-section taken perpendicular to the drive shaft.

FIG. 68 is a cross-sectional view of the rotor and drive shaft in FIG.67, with the cross-section taken along the line 68-68 in FIG. 67.

FIG. 69 is a partial cross-sectional view of a compressor according toan alternative embodiment, with the cross-section taken along an axis ofthe compressor's drive shaft.

FIG. 70 is a side view of a compressor according to an alternativeembodiment;

FIG. 71 is an end view of the compressor in FIG. 70;

FIG. 72 is a perspective side view of the compressor in FIG. 70;

FIG. 73 is a cross-sectional view of the compressor in FIG. 70, takenalong the line 73-73 in FIG. 70; and

FIG. 74 is a partial, magnified cross-sectional view of FIG. 73.

DETAILED DESCRIPTION OF THE EMBODIMENTS

To the extent that the following terms are utilized herein, thefollowing definitions are applicable:

Balanced rotation: the center of mass of the rotating mass is located onthe axis of rotation.

Chamber volume: any volume that can contain fluids for compression.

Compressor: a device used to increase the pressure of a compressiblefluid. The fluid can be either gas or vapor, and can have a widemolecular weight range.

Concentric: the center or axis of one object coincides with the centeror axis of a second object

Concentric rotation: rotation in which one object's center of rotationis located on the same axis as the second object's center of rotation.

Positive displacement compressor: a compressor that collects a fixedvolume of gas within a chamber and compresses it by reducing the chambervolume.

Proximate: sufficiently close to restrict fluid flow between highpressure and low pressure regions. Restriction does not need to beabsolute; some leakage is acceptable.

Rotor: A rotating element driven by a mechanical force to rotate aboutan axis. As used in a compressor design, the rotor imparts energy to afluid.

Rotary compressor: A positive-displacement compressor that impartsenergy to the gas being compressed by way of an input shaft moving asingle or multiple rotating elements

FIGS. 1 through 7 show external views of an embodiment of the presentinvention in which a rotary compressor includes spring backed cam drivegate positioning system. Main housing 100 includes a main casing 110 andend plates 120, each of which includes a hole through which drive shaft140 passes axially. Liquid injector assemblies 130 are located on holesin the main casing 110. The main casing includes a hole for the inletflange 160, and a hole for the gate casing 150.

Gate casing 150 is connected to and positioned below main casing 110 ata hole in main casing 110. The gate casing 150 is comprised of twoportions: an inlet side 152 and an outlet side 154. Other embodiments ofgate casing 150 may only consist of a single portion. As shown in FIG.28, the outlet side 154 includes outlet ports 435, which are holes whichlead to outlet valves 440. Alternatively, an outlet valve assembly maybe used.

Referring back to FIGS. 1-7, the spring-backed cam drive gatepositioning system 200 is attached to the gate casing 150 and driveshaft 140. The gate positioning system 200 moves gate 600 in conjunctionwith the rotation of rotor 500. A movable assembly includes gate struts210 and cam struts 230 connected to gate support arm 220 and bearingsupport plate 156. The bearing support plate 156 seals the gate casing150 by interfacing with the inlet and outlet sides through a boltedgasket connection. Bearing support plate 156 is shaped to seal gatecasing 150, mount bearing housings 270 in a sufficiently parallelmanner, and constrain compressive springs 280. In one embodiment, theinterior of the gate casing 150 is hermetically sealed by the bearingsupport plate 156 with o-rings, gaskets, or other sealing materials.Other embodiments may support the bearings at other locations, in whichcase an alternate plate may be used to seal the interior of the gatecasing. Shaft seals, mechanical seals, or other sealing mechanisms maybe used to seal around the gate struts 210 which penetrate the bearingsupport plate 156 or other sealing plate. Bearing housings 270, alsoknown as pillow blocks, are concentric to the gate struts 210 and thecam struts 230.

In the illustrated embodiment, the compressing structure comprises arotor 500. However, according to alternative embodiments, alternativetypes of compressing structures (e.g., gears, screws, pistons, etc.) maybe used in connection with the compression chamber to providealternative compressors according to alternative embodiments of theinvention.

Two cam followers 250 are located tangentially to each cam 240,providing a downward force on the gate. Drive shaft 140 turns cams 240,which transmits force to the cam followers 250. The cam followers 250may be mounted on a through shaft, which is supported on both ends, orcantilevered and only supported on one end. The cam followers 250 areattached to cam follower supports 260, which transfer the force into thecam struts 230. As cams 240 turn, the cam followers 250 are pushed down,thus moving the cam struts 230 down. This moves the gate support arm 220and the gate strut 210 down. This, in turn, moves the gate 600 down.

Springs 280 provide a restorative upward force to keep the gate 600timed appropriately to seal against the rotor 500. As the cams 240continue to turn and no longer effectuate a downward force on the camfollowers 250, springs 280 provide an upward force. As shown in thisembodiment, compression springs are utilized. As one of ordinary skillin the art would appreciate, tension springs and the shape of thebearing support plate 156 may be altered to provide for the desiredupward or downward force. The upward force of the springs 280 pushes thecam follower support 260 and thus the gate support arm 220 up which inturn moves the gate 600 up.

Due to the varying pressure angle between the cam followers 250 and cams240, the preferred embodiment may utilize an exterior cam profile thatdiffers from the rotor 500 profile. This variation in profile allows forcompensation for the changing pressure angle to ensure that the tip ofthe gate 600 remains proximate to the rotor 500 throughout the entirecompression cycle.

Line A in FIGS. 3, 6, and 7 shows the location for the cross-sectionalview of the compressor in FIG. 8. As shown in FIG. 8, the main casing110 has a cylindrical shape. Liquid injector housings 132 are attachedto, or may be cast as a part of, the main casing 110 to provide foropenings in the rotor casing 400. Because it is cylindrically shaped inthis embodiment, the rotor casing 400 may also be referenced as thecylinder. The interior wall defines a rotor casing volume 410 (alsoreferred to as the compression chamber). The rotor 500 concentricallyrotates with drive shaft 140 and is affixed to the drive shaft 140 byway of key 540 and press fit. Alternate methods for affixing the rotor500 to the drive shaft 140, such as polygons, splines, or a taperedshaft may also be used.

FIG. 9 shows an embodiment of the present invention in which a timingbelt with spring gate positioning system is utilized. This embodiment290 incorporates two timing belts 292 each of which is attached to thedrive shaft 140 by way of sheaves 294. The timing belts 292 are attachedto secondary shafts 142 by way of sheaves 295. Gate strut springs 296are mounted around gate struts. Rocker arms 297 are mounted to rockerarm supports 299. The sheaves 295 are connected to rocker arm cams 293to push the rocker arms 297 down. As the inner rings push down on oneside of the rocker arms 297, the other side pushes up against the gatesupport bar 298. The gate support bar 298 pushes up against the gatestruts and gate strut springs 296. This moves the gate up. The springs296 provide a downward force pushing the gate down.

FIGS. 10 through 17 show external views of a rotary compressorembodiment utilizing a dual cam follower gate positioning system. Themain housing 100 includes a main casing 110 and end plates 120, each ofwhich includes a hole through which a drive shaft 140 passes axially.Liquid injector assemblies 130 are located on holes in the main casing110. The main casing 110 also includes a hole for the inlet flange 160and a hole for the gate casing 150. The gate casing 150 is mounted toand positioned below the main casing 110 as discussed above.

A dual cam follower gate positioning system 300 is attached to the gatecasing 150 and drive shaft 140. The dual cam follower gate positioningsystem 300 moves the gate 600 in conjunction with the rotation of therotor 500. In a preferred embodiment, the size and shape of the cams isnearly identical to the rotor in cross-sectional size and shape. Inother embodiments, the rotor, cam shape, curvature, cam thickness, andvariations in the thickness of the lip of the cam may be adjusted toaccount for variations in the attack angle of the cam follower. Further,large or smaller cam sizes may be used. For example, a similar shape butsmaller size cam may be used to reduce roller speeds.

A movable assembly includes gate struts 210 and cam struts 230 connectedto gate support arm 220 and bearing support plate 156. In thisembodiment, the bearing support plate 157 is straight. As one ofordinary skill in the art would appreciate, the bearing support platecan utilize different geometries, including structures designed to ornot to perform sealing of the gate casing 150. In this embodiment, thebearing support plate 157 serves to seal the bottom of the gate casing150 through a bolted gasket connection. Bearing housings 270, also knownas pillow blocks, are mounted to bearing support plate 157 and areconcentric to the gate struts 210 and the cam struts 230. In certainembodiments, the components comprising this movable assembly may beoptimized to reduce weight, thereby reducing the force necessary toachieve the necessary acceleration to keep the tip of gate 600 proximateto the rotor 500. Weight reduction could additionally and/oralternatively be achieved by removing material from the exterior of anyof the moving components, as well as by hollowing out moving components,such as the gate struts 210 or the gate 600.

Drive shaft 140 turns cams 240, which transmit force to the camfollowers 250, including upper cam followers 252 and lower cam followers254. The cam followers 250 may be mounted on a through shaft, which issupported on both ends, or cantilevered and only supported on one end.In this embodiment, four cam followers 250 are used for each cam 240.Two lower cam followers 252 are located below and follow the outsideedge of the cam 240. They are mounted using a through shaft. Two uppercam followers 254 are located above the previous two and follow theinside edge of the cams 240. They are mounted using a cantileveredconnection.

The cam followers 250 are attached to cam follower supports 260, whichtransfer the force into the cam struts 230. As the cams 240 turn, thecam struts 230 move up and down. This moves the gate support arm 220 andgate struts 210 up and down, which in turn, moves the gate 600 up anddown.

Line A in FIGS. 11, 12, 15, and 16 show the location for thecross-sectional view of the compressor in FIG. 17. As shown in FIG. 17,the main casing 110 has a cylindrical shape. Liquid injector housings132 are attached to or may be cast as a part of the main casing 110 toprovide for openings in the rotor casing 400. The rotor 500concentrically rotates around drive shaft 140.

An embodiment using a belt driven system 310 is shown in FIG. 18. Timingbelts 292 are connected to the drive shaft 140 by way of sheaves 294.The timing belts 292 are each also connected to secondary shafts 142 byway of another set of sheaves 295. The secondary shafts 142 drive theexternal cams 240, which are placed below the gate casing 150 in thisembodiment. Sets of upper and lower cam followers 254 and 252 areapplied to the cams 240, which provide force to the movable assemblyincluding gate struts 210 and gate support arm 220. As one of ordinaryskill in the art would appreciate, belts may be replaced by chains orother materials.

An embodiment of the present invention using an offset gate guide systemis shown in FIGS. 19 through 22 and 33. Outlet of the compressed gas andinjected fluid is achieved through a ported gate system 602 comprised oftwo parts bolted together to allow for internal lightening features.Fluid passes through channels 630 in the upper portion of the gate 602and travels to the lengthwise sides to outlet through an exhaust port344 in a timed manner with relation to the angle of rotation of therotor 500 during the cycle. Discrete point spring-backed scraper seals326 provide sealing of the gate 602 in the single piece gate casing 336.Liquid injection is achieved through a variety of flat spray nozzles 322and injector nozzles 130 across a variety of liquid injector port 324locations and angles.

Reciprocating motion of the two-piece gate 602 is controlled through theuse of an offset spring-backed cam follower control system 320 toachieve gate motion in concert with rotor rotation. Single cams 342drive the gate system downwards through the transmission of force on thecam followers 250 through the cam struts 338. This results in controlledmotion of the crossarm 334, which is connected by bolts (some of whichare labeled as 328) with the two-piece gate 602. The crossarm 334mounted linear bushings 330, which reciprocate along the length of camshafts 332, control the motion of the gate 602 and the crossarm 334. Thecam shafts 332 are fixed in a precise manner to the main casing throughthe use of cam shaft support blocks 340. Compression springs 346 areutilized to provide a returning force on the crossarm 334, allowing thecam followers 250 to maintain constant rolling contact with the cams,thereby achieving controlled reciprocating motion of the two-piece gate602.

FIG. 23 shows an embodiment using a linear actuator system 350 for gatepositioning. A pair of linear actuators 352 is used to drive the gate.In this embodiment, it is not necessary to mechanically link the driveshaft to the gate as with other embodiments. The linear actuators 352are controlled so as to raise and lower the gate in accordance with therotation of the rotor. The actuators may be electronic, hydraulic,belt-driven, electromagnetic, gas-driven, variable-friction, or othermeans. The actuators may be computer controlled or controlled by othermeans.

FIGS. 24A and B show a magnetic drive system 360. The gate system may bedriven, or controlled, in a reciprocating motion through the placementof magnetic field generators, whether they are permanent magnets orelectromagnets, on any combination of the rotor 500, gate 600, and/orgate casing 150. The purpose of this system is to maintain a constantdistance from the tip of the gate 600 to the surface of the rotor 500 atall angles throughout the cycle. In a preferred magnetic systemembodiment, permanent magnets 366 are mounted into the ends of the rotor500 and retained. In addition, permanent magnets 364 are installed andretained in the gate 600. Poles of the magnets are aligned so that themagnetic force generated between the rotor's magnets 366 and the gate'smagnets 364 is a repulsive force, forcing the gate 600 down throughoutthe cycle to control its motion and maintain constant distance. Toprovide an upward, returning force on the gate 600, additional magnets(not shown) are installed into the bottom of the gate 600 and the bottomof the gate casing 150 to provide an additional repulsive force. Themagnetic drive systems are balanced to precisely control the gate'sreciprocating motion.

Alternative embodiments may use an alternate pole orientation to provideattractive forces between the gate and rotor on the top portion of thegate and attractive forces between the gate and gate casing on thebottom portion of the gate. In place of the lower magnet system, springsmay be used to provide a repulsive force. In each embodiment,electromagnets may be used in place of permanent magnets. In addition,switched reluctance electromagnets may also be utilized. In anotherembodiment, electromagnets may be used only in the rotor and gate. Theirpoles may switch at each inflection point of the gate's travel duringits reciprocating cycle, allowing them to be used in an attractive andrepulsive method.

Alternatively, direct hydraulic or indirect hydraulic (hydropneumatic)can be used to apply motive force/energy to the gate to drive it andposition it adequately. Solenoid or other flow control valves can beused to feed and regulate the position and movement of the hydraulic orhydropneumatic elements. Hydraulic force may be converted to mechanicalforce acting on the gate through the use of a cylinder based or directhydraulic actuators using membranes/diaphragms.

FIG. 25 shows an embodiment using a scotch yoke gate positioning system370. Here, a pair of scotch yokes 372 is connected to the drive shaftand the bearing support plate. A roller rotates at a fixed radius withrespect to the shaft. The roller follows a slot within the yoke 372,which is constrained to a reciprocating motion. The yoke geometry can bemanipulated to a specific shape that will result in desired gatedynamics.

As one of skill in the art would appreciate, these alternative drivemechanisms do not require any particular number of linkages between thedrive shaft and the gate. For example, a single spring, belt, linkagebar, or yoke could be used. Depending on the design implementation, morethan two such elements could be used.

FIGS. 26A-26F show a compression cycle of an embodiment utilizing a tipseal 620. As the drive shaft 140 turns, the rotor 500 and gate strut 210push up gate 600 so that it is timed with the rotor 500. As the rotor500 turns clockwise, the gate 600 rises up until the rotor 500 is in the12 o'clock position shown in FIG. 26C. As the rotor 500 continues toturn, the gate 600 moves downward until it is back at the 6 o'clockposition in FIG. 26F. The gate 600 separates the portion of the cylinderthat is not taken up by rotor 500 into two components: an intakecomponent 412 and a compression component 414. In one embodiment, tipseal 620 may not be centered within the gate 600, but may instead beshifted towards one side so as to minimize the area on the top of thegate on which pressure may exert a downwards force on the gate. This mayalso have the effect of minimizing the clearance volume of the system.In another embodiment, the end of the tip seal 620 proximate to therotor 500 may be rounded, so as to accommodate the varying contact anglethat will be encountered as the tip seal 620 contacts the rotor 500 atdifferent points in its rotation.

FIGS. 26A-F depict steady state operation. Accordingly, in FIG. 26A,where the rotor 500 is in the 6 o'clock position, the compression volume414, which constitutes a subset of the rotor casing volume 410, alreadyhas received fluid. In FIG. 26B, the rotor 500 has turned clockwise andgate 600 has risen so that the tip seal 620 makes contact with the rotor500 to separate the intake volume 412, which also constitutes a subsetof the rotor casing volume 410, from the compression volume 414.Embodiments using the roller tip 650 discussed below instead of tip seal620 would operate similarly. As the rotor 500 turns, as shown further inFIGS. 26C-E, the intake volume 412 increases, thereby drawing in morefluid from inlet 420, while the compression volume 414 decreases. As thevolume of the compression volume 414 decreases, the pressure increases.The pressurized fluid is then expelled by way of an outlet 430. At apoint in the compression cycle when a desired high pressure is reached,the outlet valve opens and the high pressure fluid can leave thecompression volume 414. In this embodiment, the valve outputs both thecompressed gas and the liquid injected into the compression chamber.

FIGS. 27A-27F show an embodiment in which the gate 600 does not use atip seal. Instead, the gate 600 is timed to be proximate to the rotor500 as it turns. The close proximity of the gate 600 to the rotor 500leaves only a very small path for high pressure fluid to escape. Closeproximity in conjunction with the presence of liquid (due to the liquidinjectors 136 or an injector placed in the gate itself) allow the gate600 to effectively create an intake fluid component 412 and acompression component 414. Embodiments incorporating notches 640 wouldoperate similarly.

FIG. 28 shows a cross-sectional perspective view of the rotor casing400, the rotor 500, and the gate 600. The inlet port 420 shows the paththat gas can enter. The outlet 430 is comprised of several holes thatserve as outlet ports 435 that lead to outlet valves 440. The gatecasing 150 consists of an inlet side 152 and an outlet side 154. Areturn pressure path (not shown) may be connected to the inlet side 152of the gate casing 150 and the inlet port 420 to ensure that there is noback pressure build up against gate 600 due to leakage through the gateseals. As one of ordinary skill in the art would appreciate, it isdesirable to achieve a hermetic seal, although perfect hermetic sealingis not necessary.

In alternate embodiments, the outlet ports 435 may be located in therotor casing 400 instead of the gate casing 150. They may be located ata variety of different locations within the rotor casing. The outletvalves 440 may be located closer to the compression chamber, effectivelyminimizing the volume of the outlet ports 430, to minimize the clearancevolume related to these outlet ports. A valve cartridge may be usedwhich houses one or more outlet valves 440 and connects directly to therotor casing 400 or gate casing 150 to align the outlet valves 440 withoutlet ports 435. This may allow for ease of installing and removing theoutlet valves 440.

FIG. 29 shows an alternative embodiment in which flat spray liquidinjector housings 170 are located on the main casing 110 atapproximately the 3 o'clock position. These injectors can be used toinject liquid directly onto the inlet side of the gate 600, ensuringthat it does not reach high temperatures. These injectors also help toprovide a coating of liquid on the rotor 500, helping to seal thecompressor.

As discussed above, the preferred embodiments utilize a rotor thatconcentrically rotates within a rotor casing. In the preferredembodiment, the rotor 500 is a right cylinder with a non-circularcross-section that runs the length of the main casing 110. FIG. 30 showsa cross-sectional view of the sealing and non-sealing portions of therotor 500. The profile of the rotor 500 is comprised of three sections.The radii in sections I and III are defined by a cycloidal curve. Thiscurve also represents the rise and fall of the gate and defines anoptimum acceleration profile for the gate. Other embodiments may usedifferent curve functions to define the radius such as a double harmonicfunction. Section II employs a constant radius 570, which corresponds tothe maximum radius of the rotor. The minimum radius 580 is located atthe intersection of sections I and III, at the bottom of rotor 500. In apreferred embodiment, D is 23.8 degrees. In alternative embodiments,other angles may be utilized depending on the desired size of thecompressor, the desired acceleration of the gate, and desired sealingarea.

The radii of the rotor 500 in one preferred embodiment can be calculatedusing the following functions:

${r(t)} = \left\{ \begin{matrix}{r_{I} = {r_{\min} + {h\left\lbrack {\frac{t_{I}}{T} + {\sin \left( \frac{2\; \pi \; t_{I}}{T} \right)}} \right\rbrack}}} \\{r_{II} = r_{\max}} \\{r_{III} = {r_{\min} + {h\left\lbrack {\frac{t_{III}}{T} + {\sin \left( \frac{2\; \pi \; t_{III}}{T} \right)}} \right\rbrack}}}\end{matrix} \right.$

According to an alternative embodiment, the radii of the rotor 500 iscalculated as a 3-4-5-polynomial function.

In a preferred embodiment, the rotor 500 is symmetrical along one axis.It may generally resemble a cross-sectional egg shape. The rotor 500includes a hole 530 in which the drive shaft 140 and a key 540 may bemounted. The rotor 500 has a sealing section 510, which is the outersurface of the rotor 500 corresponding to section II, and a non-sealingsection 520, which is the outer surface of the rotor 500 correspondingto sections I and III. The sections I and III have a smaller radius thansections II creating a compression volume. The sealing portion 510 isshaped to correspond to the curvature of the rotor casing 400, therebycreating a dwell seal that effectively minimizes communication betweenthe outlet 430 and inlet 420. Physical contact is not required for thedwell seal. Instead, it is sufficient to create a tortuous path thatminimizes the amount of fluid that can pass through. In a preferredembodiment, the gap between the rotor and the casing in this embodimentis less than 0.008 inches. As one of ordinary skill in the art wouldappreciate, this gap may be altered depending on tolerances, both inmachining the rotor 500 and rotor housing 400, temperature, materialproperties, and other specific application requirements.

Additionally, as discussed below, liquid is injected into thecompression chamber. By becoming entrained in the gap between thesealing portion 510 and the rotor casing 400, the liquid can increasethe effectiveness of the dwell seal.

As shown in FIG. 31A, the rotor 500 is balanced with cut out shapes andcounterweights. Holes, some of which are marked as 550, lighten therotor 500. These lightening holes may be filled with a low densitymaterial to ensure that liquid cannot encroach into the rotor interior.Alternatively, caps may be placed on the ends of rotor 500 to seal thelightening holes. Counterweights, one of which is labeled as 560, aremade of a denser material than the remainder of the rotor 500. Theshapes of the counterweights can vary and do not need to be cylindrical.

The rotor design provides several advantages. As shown in the embodimentof FIG. 31A, the rotor 500 includes 7 cutout holes 550 on one side andtwo counterweights 560 on the other side to allow the center of mass tomatch the center of rotation. An opening 530 includes space for thedrive shaft and a key. This weight distribution is designed to achievebalanced, concentric motion. The number and location of cutouts andcounterweights may be changed depending on structural integrity, weightdistribution, and balanced rotation parameters. In various embodiments,cutouts and/or counterweights or neither may be used required to achievebalanced rotor rotation.

The cross-sectional shape of the rotor 500 allows for concentricrotation about the drive shaft's axis of rotation, a dwell seal 510portion, and open space on the non-sealing side for increased gas volumefor compression. Concentric rotation provides for rotation about thedrive shaft's principal axis of rotation and thus smoother motion andreduced noise.

An alternative rotor design 502 is shown in FIG. 31B. In thisembodiment, a different arc of curvature is implemented utilizing threeholes 550 and a circular opening 530. Another alternative design 504 isshown in FIG. 31C. Here, a solid rotor shape is used and a larger hole530 (for a larger drive shaft) is implemented. Yet another alternativerotor design 506 is shown in FIG. 31D incorporating an asymmetricalshape, which would smooth the volume reduction curve, allowing forincreased time for heat transfer to occur at higher pressures.Alternative rotor shapes may be implemented for different curvatures orneeds for increased volume in the compression chamber.

The rotor surface may be smooth in embodiments with contacting tip sealsto minimize wear on the tip seal. In alternative embodiments, it may beadvantageous to put surface texture on the rotor to create turbulencethat may improve the performance of non-contacting seals. In otherembodiments, the rotor casing's interior cylindrical wall may further betextured to produce additional turbulence, both for sealing and heattransfer benefits. This texturing could be achieved through machining ofthe parts or by utilizing a surface coating. Another method of achievingthe texture would be through blasting with a waterjet, sandblast, orsimilar device to create an irregular surface.

The main casing 110 may further utilize a removable cylinder liner. Thisliner may feature microsurfacing to induce turbulence for the benefitsnoted above. The liner may also act as a wear surface to increase thereliability of the rotor and casing. The removable liner could bereplaced at regular intervals as part of a recommended maintenanceschedule. The rotor may also include a liner. Sacrifical or wear-incoatings may be used on the rotor 500 or rotor casing 400 to correct formanufacturing defects in ensuring the preferred gap is maintained alongthe sealing portion 510 of the rotor 500.

The exterior of the main casing 110 may also be modified to meetapplication specific parameters. For example, in subsea applications,the casing may require to be significantly thickened to withstandexterior pressure, or placed within a secondary pressure vessel. Otherapplications may benefit from the exterior of the casing having arectangular or square profile to facilitate mounting exterior objects orstacking multiple compressors. Liquid may be circulated in the casinginterior to achieve additional heat transfer or to equalize pressure inthe case of subsea applications for example.

As shown in FIGS. 32A and B, the combination of the rotor 500 (heredepicted with rotor end caps 590), the gate 600, and drive shaft 140,provide for a more efficient manner of compressing fluids in a cylinder.The gate is aligned along the length of the rotor to separate and definethe inlet portion and compression portion as the rotor turns.

The drive shaft 140 is mounted to endplates 120 in the preferredembodiment using one spherical roller bearing in each endplate 120. Morethan one bearing may be used in each endplate 120, in order to increasetotal load capacity. A grease pump (not shown) is used to providelubrication to the bearings. Various types of other bearings may beutilized depending on application specific parameters, including rollerbearings, ball bearings, needle bearings, conical bearings, cylindricalbearings, journal bearings, etc. Different lubrication systems usinggrease, oil, or other lubricants may also be used. Further, drylubrication systems or materials may be used. Additionally, applicationsin which dynamic imbalance may occur may benefit from multi-bearingarrangements to support stray axial loads.

Operation of gates in accordance with embodiments of the presentinvention are shown in FIGS. 8, 17, 22, 24B, 26A-F, 27A-F, 28, 32A-B,and 33-36. As shown in FIGS. 26A-F and 27A-F, gate 600 creates apressure boundary between an intake volume 412 and a compression volume414. The intake volume 412 is in communication with the inlet 420. Thecompression volume 414 is in communication with the outlet 430.Resembling a reciprocating, rectangular piston, the gate 600 rises andfalls in time with the turning of the rotor 500.

The gate 600 may include an optional tip seal 620 that makes contactwith the rotor 500, providing an interface between the rotor 500 and thegate 600. Tip seal 620 consists of a strip of material at the tip of thegate 600 that rides against rotor 500. The tip seal 620 could be made ofdifferent materials, including polymers, graphite, and metal, and couldtake a variety of geometries, such as a curved, flat, or angled surface.The tip seal 620 may be backed by pressurized fluid or a spring forceprovided by springs or elastomers. This provides a return force to keepthe tip seal 620 in sealing contact with the rotor 500.

Different types of contacting tips may be used with the gate 600. Asshown in FIG. 35, a roller tip 650 may be used. The roller tip 650rotates as it makes contact with the turning rotor 500. Also, tips ofdiffering strengths may be used. For example, a tip seal 620 or rollertip 650 may be made of softer metal that would gradually wear downbefore the rotor 500 surfaces would wear.

Alternatively, a non-contacting seal may be used. Accordingly, the tipseal may be omitted. In these embodiments, the topmost portion of thegate 600 is placed proximate, but not necessarily in contact with, therotor 500 as it turns. The amount of allowable gap may be adjusteddepending on application parameters.

As shown in FIGS. 34A and 34B, in an embodiment in which the tip of thegate 600 does not contact the rotor 500, the tip may include notches 640that serve to keep gas pocketed against the tip of the gate 600. Theentrained fluid, in either gas or liquid form, assists in providing anon-contacting seal. As one of ordinary skill in the art wouldappreciate, the number and size of the notches is a matter of designchoice dependent on the compressor specifications.

Alternatively, liquid may be injected from the gate itself. As shown inFIG. 36, a cross-sectional view of a portion of a gate, one or morechannels 660 from which a fluid may pass may be built into the gate. Inone such embodiment, a liquid can pass through a plurality of channels660 to form a liquid seal between the topmost portion of the gate 600and the rotor 500 as it turns. In another embodiment, residualcompressed fluid may be inserted through one or more channels 660.Further still, the gate 600 may be shaped to match the curvature ofportions of the rotor 500 to minimize the gap between the gate 600 andthe rotor 500.

Preferred embodiments enclose the gate in a gate casing. As shown inFIGS. 8 and 17, the gate 600 is encompassed by the gate casing 150,including notches, one of which is shown as item 158. The notches holdthe gate seals, which ensure that the compressed fluid will not releasefrom the compression volume 414 through the interface between gate 600and gate casing 150 as gate 600 moves up and down. The gate seals may bemade of various materials, including polymers, graphite or metal. Avariety of different geometries may be used for these seals. Variousembodiments could utilize different notch geometries, including ones inwhich the notches may pass through the gate casing, in part or in full.

In alternate embodiments, the seals could be placed on the gate 600instead of within the gate casing 150. The seals would form a ringaround the gate 600 and move with the gate relative to the casing 150,maintaining a seal against the interior of the gate casing 150. Thelocation of the seals may be chosen such that the center of pressure onthe gate 600 is located on the portion of the gate 600 inside of thegate casing 150, thus reducing or eliminating the effect of acantilevered force on the portion of the gate 600 extending into therotor casing 400. This may help eliminate a line contact between thegate 600 and gate casing 150 and instead provide a surface contact,allowing for reduced friction and wear. One or more wear plates may beused on the gate 600 to contact the gate casing 150. The location of theseals and wear plates may be optimized to ensure proper distribution offorces across the wear plates.

The seals may use energizing forces provided by springs or elastomerswith the assembly of the gate casing 150 inducing compression on theseals. Pressurized fluid may also be used to energize the seals.

The gate 600 is shown with gate struts 210 connected to the end of thegate. In various embodiments, the gate 600 may be hollowed out such thatthe gate struts 210 can connect to the gate 600 closer to its tip. Thismay reduce the amount of thermal expansion encountered in the gate 600.A hollow gate also reduces the weight of the moving assembly and allowsoil or other lubricants and coolants to be splashed into the interior ofthe gate to maintain a cooler temperature. The relative location ofwhere the gate struts 210 connect to the gate 600 and where the gateseals are located may be optimized such that the deflection modes of thegate 600 and gate struts 210 are equal, allowing the gate 600 to remainparallel to the interior wall of the gate casing 150 when it deflectsdue to pressure, as opposed to rotating from the pressure force.Remaining parallel may help to distribute the load between the gate 600and gate casing 150 to reduce friction and wear.

A rotor face seal may also be placed on the rotor 500 to provide for aninterface between the rotor 500 and the endplates 120. An outer rotorface seal is placed along the exterior edge of the rotor 500, preventingfluid from escaping past the end of the rotor 500. A secondary innerrotor face seal is placed on the rotor face at a smaller radius toprevent any fluid that escapes past the outer rotor face seal fromescaping the compressor entirely. This seal may use the same or othermaterials as the gate seal. Various geometries may be used to optimizethe effectiveness of the seals. These seals may use energizing forcesprovided by springs, elastomers or pressurized fluid. Lubrication may beprovided to these rotor face seals by injecting oil or other lubricantthrough ports in the endplates 120.

Along with the seals discussed herein, the surfaces those seals contact,known as counter-surfaces, may also be considered. In variousembodiments, the surface finish of the counter-surface may besufficiently smooth to minimize friction and wear between the surfaces.In other embodiments, the surface finish may be roughened or given apattern such as cross-hatching to promote retention of lubricant orturbulence of leaking fluids. The counter-surface may be composed of aharder material than the seal to ensure the seal wears faster than thecounter-surface, or the seal may be composed of a harder material thanthe counter-surface to ensure the counter-surface wears faster than theseal. The desired physical properties of the counter-surface (surfaceroughness, hardness, etc.) may be achieved through material selection,material finishing techniques such as quenching, tempering, or workhardening, or selection and application of coatings that achieve thedesired characteristics. Final manufacturing processes, such as surfacegrinding, may be performed before or after coatings are applied. Invarious embodiments, the counter-surface material may be steel orstainless steel. The material may be hardened via quenching ortempering. A coating may be applied, which could be chrome, titaniumnitride, silicon carbide, or other materials.

Minimizing the possibility of fluids leaking to the exterior of the mainhousing 100 is desirable. Various seals, such as gaskets and o-rings,are used to seal external connections between parts. For example, in apreferred embodiment, a double o-ring seal is used between the maincasing 110 and endplates 120. Further seals are utilized around thedrive shaft 140 to prevent leakage of any fluids making it past therotor face seals. A lip seal is used to seal the drive shaft 140 whereit passes through the endplates 120. In various embodiments, multipleseals may be used along the drive shaft 140 with small gaps between themto locate vent lines and hydraulic packings to reduce or eliminate gasleakage exterior to the compression chamber. Other forms of seals couldalso be used, such as mechanical or labyrinth seals.

It is desirable to achieve near isothermal compression. To providecooling during the compression process, liquid injection is used. Inpreferred embodiments, the liquid is atomized to provide increasedsurface area for heat absorption. In other embodiments, different sprayapplications or other means of injecting liquids may be used.

Liquid injection is used to cool the fluid as it is compressed,increasing the efficiency of the compression process. Cooling allowsmost of the input energy to be used for compression rather than heatgeneration in the gas. The liquid has dramatically superior heatabsorption characteristics compared to gas, allowing the liquid toabsorb heat and minimize temperature increase of the working fluid,achieving near isothermal compression. As shown in FIGS. 8 and 17,liquid injector assemblies 130 are attached to the main casing 110.Liquid injector housings 132 include an adapter for the liquid source134 (if it is not included with the nozzle) and a nozzle 136. Liquid isinjected by way of a nozzle 136 directly into the rotor casing volume410.

The amount and timing of liquid injection may be controlled by a varietyof implements including a computer-based controller capable of measuringthe liquid drainage rate, liquid levels in the chamber, and/or anyrotational resistance due to liquid accumulation through a variety ofsensors. Valves or solenoids may be used in conjunction with the nozzlesto selectively control injection timing. Variable orifice control mayalso be used to regulate the amount of liquid injection and othercharacteristics.

Analytical and experimental results are used to optimize the number,location, and spray direction of the injectors 136. These injectors 136may be located in the periphery of the cylinder. Liquid injection mayalso occur through the rotor or gate. The current embodiment of thedesign has two nozzles located at 12 o'clock and 10 o'clock. Differentapplication parameters will also influence preferred nozzle arrays.

Because the heat capacity of liquids is typically much higher thangases, the heat is primarily absorbed by the liquid, keeping gastemperatures lower than they would be in the absence of such liquidinjection.

When a fluid is compressed, the pressure times the volume raised to apolytropic exponent remains constant throughout the cycle, as seen inthe following equation:

P*V ^(n)=Constant

In polytropic compression, two special cases represent the opposingsides of the compression spectrum. On the high end, adiabaticcompression is defined by a polytropic constant of n=1.4 for air, orn=1.28 for methane. Adiabatic compression is characterized by thecomplete absence of cooling of the working fluid (isentropic compressionis a subset of adiabatic compression in which the process isreversible). This means that as the volume of the fluid is reduced, thepressure and temperature each rise accordingly. It is an inefficientprocess due to the exorbitant amount of energy wasted in the generationof heat in the fluid, which often needs to be cooled down again later.Despite being an inefficient process, most conventional compressiontechnology, including reciprocating piston and centrifugal typecompressors are essentially adiabatic. The other special case isisothermal compression, where n=1. It is an ideal compression cycle inwhich all heat generated in the fluid is transmitted to the environment,maintaining a constant temperature in the working fluid. Although itrepresents an unachievable perfect case, isothermal compression isuseful in that it provides a lower limit to the amount of energyrequired to compress a fluid.

FIG. 37 shows a sample pressure-volume (P-V) curve comparing severaldifferent compression processes. The isothermal curve shows thetheoretically ideal process. The adiabatic curve represents an adiabaticcompression cycle, which is what most conventional compressortechnologies follow. Since the area under the P-V curve represents theamount of work required for compression, approaching the isothermalcurve means that less work is needed for compression. A model of one ormore compressors according to various embodiments of the presentinvention is also shown, nearly achieving as good of results as theisothermal process. According to various embodiments, theabove-discussed coolant injection facilitates the near isothermalcompression through absorption of heat by the coolant. Not only doesthis near-isothermal compression process require less energy, at the endof the cycle gas temperatures are much lower than those encountered withtraditional compressors. According to various embodiments, such areduction in compressed working fluid temperature eliminates the use ofor reduces the size of expensive and efficiency-robbing after-coolers.

Embodiments of the present invention achieve these near-isothermalresults through the above-discussed injection of liquid coolant.Compression efficiency is improved according to one or more embodimentsbecause the working fluid is cooled by injecting liquid directly intothe chamber during the compression cycle. According to variousembodiments, the liquid is injected directly into the area of thecompression chamber where the gas is undergoing compression.

Rapid heat transfer between the working fluid and the coolant directlyat the point of compression may facilitate high pressure ratios. Thatleads to several aspects of various embodiments of the present inventionthat may be modified to improve the heat transfer and raise the pressureratio.

One consideration is the heat capacity of the liquid coolant. The basicheat transfer equation is as follows:

Q=mc _(p) ΔT

-   -   where Q is the heat, m is mass, ΔT is change in temperature, and        c_(p) is the specific heat.        The higher the specific heat of the coolant, the more heat        transfer that will occur.

Choosing a coolant is sometimes more complicated than simply choosing aliquid with the highest heat capacity possible. Other factors, such ascost, availability, toxicity, compatibility with working fluid, andothers can also be considered. In addition, other characteristics of thefluid, such as viscosity, density, and surface tension affect thingslike droplet formation which, as will be discussed below, also affectcooling performance.

According to various embodiments, water is used as the cooling liquidfor air compression. For methane compression, various liquidhydrocarbons may be effective coolants, as well as triethylene glycol.

Another consideration is the relative velocity of coolant to the workingfluid. Movement of the coolant relative to the working fluid at thelocation of compression of the working fluid (which is the point of heatgeneration) enhances heat transfer from the working fluid to thecoolant. For example, injecting coolant at the inlet of a compressorsuch that the coolant is moving with the working fluid by the timecompression occurs and heat is generated will cool less effectively thanif the coolant is injected in a direction perpendicular to or counter tothe flow of the working fluid adjacent the location of liquid coolantinjection. FIGS. 38(a)-(d) show a schematic of the sequentialcompression cycle in a compressor according to an embodiment of theinvention. The dotted arrows in FIG. 38(c) show the injection locations,directions, and timing used according to various embodiments of thepresent invention to enhance the cooling performance of the system.

As shown in FIG. 38(a), the compression stroke begins with a maximumworking fluid volume (shown in gray) within the compression chamber. Inthe illustrated embodiment, the beginning of the compression strokeoccurs when the rotor is at the 6 o'clock position (in an embodiment inwhich the gate is disposed at 6 o'clock with the inlet on the left ofthe gate and the outlet on the right of the gate as shown in FIGS.38(a)-(d)). In FIG. 38(b), compression has started, the rotor is at the9 o'clock position, and cooling liquid is injected into the compressionchamber. In FIG. 38(c), about 50% of the compression stroke hasoccurred, and the rotor is disposed at the 12 o'clock position. FIG.38(d) illustrates a position (3 o'clock) in which the compression strokeis nearly completed (e.g., about 95% complete). Compression isultimately completed when the rotor returns to the position shown inFIG. 38(a).

As shown in FIGS. 38(b) and (c), dotted arrows illustrate the timing,location, and direction of the coolant injection.

According to various embodiments, coolant injection occurs during onlypart of the compression cycle. For example, in each compressioncycle/stroke, the coolant injection may begin at or after the first 10,20, 30, 40, 50, 60 and/or 70% of the compression stroke/cycle (thestroke/cycle being measured in terms of volumetric compression).According to various embodiments, the coolant injection may end at eachnozzle shortly before the rotor sweeps past the nozzle (e.g., resultingin sequential ending of the injection at each nozzle (clockwise asillustrated in FIG. 38)). According to various alternative embodiments,coolant injection occurs continuously throughout the compression cycle,regardless of the rotor position.

As shown in FIGS. 38(b) and (c), the nozzles inject the liquid coolantinto the chamber perpendicular to the sweeping direction of the rotor(i.e., toward the rotor's axis of rotation, in the inward radialdirection relative to the rotor's axis of rotation). However, accordingto alternative embodiments, the direction of injection may be orientedso as to aim more upstream (e.g., at an acute angle relative to theradial direction such that the coolant is injected in a partiallycounter-flow direction relative to the sweeping direction of the rotor).According to various embodiments, the acute angle may be anywherebetween 0 and 90 degrees toward the upstream direction relative to theradial line extending from the rotor's axis of rotation to the injectornozzle. Such an acute angle may further increase the velocity of thecoolant relative to the surrounding working fluid, thereby furtherenhancing the heat transfer.

A further consideration is the location of the coolant injection, whichis defined by the location at which the nozzles inject coolant into thecompression chamber. As shown in FIGS. 38(b) and (c), coolant injectionnozzles are disposed at about 1, 2, 3, and 4 o'clock. However,additional and/or alternative locations may be chosen without deviatingfrom the scope of the present invention. According to variousembodiments, the location of injection is positioned within thecompression volume (shown in gray in FIG. 38) that exists during thecompressor's highest rate of compression (in terms of Δvolume/time orΔvolume/degree-of-rotor-rotation, which may or may not coincide). In theembodiment illustrated in FIG. 38, the highest rate of compressionoccurs around where the rotor is rotating from the 12 o'clock positionshown in FIG. 38(c) to the 3 o'clock position shown in FIG. 38(d). Thislocation is dependent on the compression mechanism being employed and invarious embodiments of the invention may vary. An injection location mayalso be selected at an earlier location in the compression chamber (e.g.9 o'clock in FIGS. 38 (a)-(d) to minimize the pressure against which theliquid must be injected, thus reducing the power required for coolantinjection. Additionally and/or alternatively, liquid (e.g., coolant) maybe injected into the inlet port before the working fluid reaches thecompression chamber.

As one skilled in the art could appreciate, the number and location ofthe nozzles may be selected based on a variety of factors. The number ofnozzles may be as few as 1 or as many as 256 or more. According tovarious embodiments, the compressor includes (a) at least 1, 2, 3, 4, 5,6, 7, 8, 9, 10, 15, 20, 30, 40, 50, 75, 100, 125, 150, 175, 200, 225,and/or 250 nozzles, (b) less than 400, 300, 275, 250, 225, 200, 175,150, 125, 100, 75, 50, 40, 30, 20, 15, and/or 10 nozzles, (c) between 1and 400 nozzles, and/or (d) any range of nozzles bounded by such numbersof any ranges therebetween. According to various embodiments, liquidcoolant injection may be avoided altogether such that no nozzles areused. Along with varying the location along the angle of the rotorcasing, a different number of nozzles may be installed at variouslocations along the length of the rotor casing. In certain embodiments,the same number of nozzles will be placed along the length of the casingat various angles. In other embodiments, nozzles may bescattered/staggered at different locations along the casing's lengthsuch that a nozzle at one angle may not have another nozzle at exactlythe same location along the length at other angles. In variousembodiments, a manifold may be used in which one or more nozzle isinstalled that connects directly to the rotor casing, simplifying theinstallation of multiple nozzles and the connection of liquid lines tothose nozzles.

Coolant droplet size is a further consideration. Because the rate ofheat transfer is linearly proportional to the surface area of liquidacross which heat transfer can occur, the creation of smaller dropletsvia the above-discussed atomizing nozzles improves cooling by increasingthe liquid surface area and allowing heat transfer to occur morequickly. Reducing the diameter of droplets of coolant in half (for agiven mass) increases the surface area by a factor of two and thusimproves the rate of heat transfer by a factor of 2. In addition, forsmall droplets the rate of convection typically far exceeds the rate ofconduction, effectively creating a constant temperature across thedroplet and removing any temperature gradients. This may result in thefull mass of liquid being used to cool the gas, as opposed to largerdroplets where some mass at the center of the droplet may not contributeto the cooling effect. Based on that evidence, it appears advantageousto inject as small of droplets as possible. However, droplets that aretoo small, when injected into the high density, high turbulence regionas shown in FIGS. 38(b) and (c), run the risk of being swept up by theworking fluid and not continuing to move through the working fluid andmaintain high relative velocity. Small droplets may also evaporate andlead to deposition of solids on the compressor's interior surfaces.Other extraneous factors also affect droplet size decisions, such aspower losses of the coolant being forced through the nozzle and amountof liquid that the compressor can handle internally.

According to various embodiments, average droplet sizes of between 50and 500 microns, between 50 and 300 microns, between 100 and 150microns, and/or any ranges within those ranges, may be fairly effective.

The mass of the coolant liquid is a further consideration. As evidencedby the heat equation shown above, more mass (which is proportional tovolume) of coolant will result in more heat transfer. However, the massof coolant injected may be balanced against the amount of liquid thatthe compressor can accommodate, as well as extraneous power lossesrequired to handle the higher mass of coolant. According to variousembodiments, between 1 and 100 gallons per minute (gpm), between 3 and40 gpm, between 5 and 25 gpm, between 7 and 10 gpm, and/or any rangestherebetween may provide an effective mass flow rate (averagedthroughout the compression stroke despite the non-continuous injectionaccording to various embodiments). According to various embodiments, thevolumetric flow rate of liquid coolant into the compression chamber maybe at least 1, 2, 3, 4, 5, 6, 7, 8, 9, and/or 10 gpm. According tovarious embodiments, flow rate of liquid coolant into the compressionchamber may be less than 100, 80, 60, 50, 40, 30, 25, 20, 15, and/or 10gpm.

The nozzle array may be designed for a high flow rate of greater than 1,2, 3, 4, 5, 6, 7, 8, 9, 10, and/or 15 gallons per minute and be capableof extremely small droplet sizes of less than 500 and/or 150 microns orless at a low differential pressure of less than 400, 300, 200, and/or100 psi. Two exemplary nozzles are Spraying Systems Co. Part Number:1/4HHSJ-SS12007 and Bex Spray Nozzles Part Number: 1/4YS12007. Othernon-limiting nozzles that may be suitable for use in various embodimentsinclude Spraying Systems Co. Part Number 1/4LN-SS14 and 1/4LN-SS8. Thepreferred flow rate and droplet size ranges will vary with applicationparameters. Alternative nozzle styles may also be used. For example, oneembodiment may use micro-perforations in the cylinder through which toinject liquid, counting on the small size of the holes to createsufficiently small droplets. Other embodiments may include various offthe shelf or custom designed nozzles which, when combined into an array,meet the injection requirements necessary for a given application.

According to various embodiments, one, several, and/or all of theabove-discussed considerations, and/or additional/alternative externalconsiderations may be balanced to optimize the compressor's performance.Although particular examples are provided, different compressor designsand applications may result in different values being selected.

According to various embodiments, the coolant injection timing,location, and/or direction, and/or other factors, and/or the higherefficiency of the compressor facilitates higher pressure ratios. As usedherein, the pressure ratio is defined by a ratio of (1) the absoluteinlet pressure of the source working fluid coming into the compressionchamber (upstream pressure) to (2) the absolute outlet pressure of thecompressed working fluid being expelled from the compression chamber(downstream pressure downstream from the outlet valve). As a result, thepressure ratio of the compressor is a function of the downstream vessel(pipeline, tank, etc.) into which the working fluid is being expelled.Compressors according to various embodiments of the present inventionwould have a 1:1 pressure ratio if the working fluid is being taken fromand expelled into the ambient environment (e.g., 14.7 psia/14.7 psia).Similarly, the pressure ratio would be about 26:1 (385 psia/14.7 psia)according to various embodiments of the invention if the working fluidis taken from ambient (14.7 psia upstream pressure) and expelled into avessel at 385 psia (downstream pressure).

According to various embodiments, the compressor has a pressure ratio of(1) at least 3:1, 4:1, 5:1, 6:1, 8:1, 10:1, 15:1, 20:1, 25:1, 30:1,35:1, and/or 40:1 or higher, (2) less than or equal to 200:1, 150:1,125:1, 100:1, 90:1, 80:1, 70:1, 60:1, 50:1, 45:1, 40:1, 35:1, and/or30:1, and (3) any and all combinations of such upper and lower ratios(e.g., between 10:1 and 200:1, between 15:1 and 100:1, between 15:1 and80:1, between 15:1 and 50:1, etc.).

According to various embodiments, lower pressure ratios (e.g., between3:1 and 15:1) may be used for working fluids with higher liquid content(e.g., with a liquid volume fraction at the compressor's inlet port ofat least 0.5, 1, 2, 3, 4, 5, 6, 7, 8, 9, 10, 15, 20, 25, 30, 35, 40, 50,60, 70, 75, 80, 85, 90, 91, 92, 93, 94, 95, 96, 97, 98, and/or 99%).Conversely, according to various embodiments, higher pressure ratios(e.g., above 15:1) may be used for working fluids with lower liquidcontent relative to gas content. However, wetter gases may nonethelessbe compressed at higher pressure ratios and drier gases may becompressed at lower pressure ratios without deviating from the scope ofvarious embodiments of the present invention.

Various embodiments of the invention are suitable for alternativeoperation using a variety of different operational parameters. Forexample, a single compressor according to one or more embodiments may besuitable to efficiently compress working fluids having drasticallydifferent liquid volume fractions and at different pressure ratios. Forexample, a compressor according to one or more embodiments is suitablefor alternatively (1) compressing a working fluid with a liquid volumefraction of between 10 and 50 percent at a pressure ratio of between 3:1and 15:1, and (2) compressing a working fluid with a liquid volumefraction of less than 10 percent at a pressure ratio of at least 15:1,20:1, 30:1, and/or 40:1.

According to various embodiments, the compressor efficiently andcost-effectively compresses both wet and dry gas using a high pressureratio.

According to various embodiments, the compressor is capable of and runsat commercially viable speeds (e.g., between 450 and 1800 rpm).According to various embodiments, the compressor runs at a speed of (a)at least 350, 400, 450, 500, 550, 600, and/or 650 rpm, (b) less than orequal to 3000, 2500, 2000, 1800, 1700, 1600, 1500, 1400, 1300, 1200,1100, 1050, 1000, 950, 900, 850, and/or 800 rpm, and/or (c) between 350and 300 rpm, 450-1800 rpm, and/or any ranges within these non-limitingupper and lower limits. According to various embodiments, the compressoris continuously operated at one or more of these speeds for at least0.5, 1, 5, 10, 15, 20, 30, 60, 90, 100, 150, 200, 250 300, 350, 400,450, and/or 500 minutes and/or at least 10, 20, 24, 48, 72, 100, 200,300, 400, and/or 500 hours.

According to various embodiments, the outlet pressure of the compressedfluid is (1) at least 200, 225, 250, 275, 300, 325, 350, 375, 400, 425,450, 475, 500, 600, 700, 800, 900, 1000, 1250, 1500, 2000, 3000, 4000,and/or 5000 psig, (2) less than 6000, 5500, 5000, 4000, 3000, 2500,2250, 2000, 1750, 1500, 1250, 1100, 1000, 900, 800, 700, 600 and/or 500psig, (3) between 200 and 6000 psig, between 200 and 5000 psig, and/or(4) within any range between the upper and lower pressures describedabove.

According to various embodiments, the inlet pressure is ambient pressurein the environment surrounding the compressor (e.g., 1 atm, 14.7 psia).Alternatively, the inlet pressure could be close to a vacuum (near 0psia), or anywhere therebetween. According to alternative embodiments,the inlet pressure may be (1) at least −14.5, −10, −5, 0, 5, 10, 25, 50,100, 150, 200, 250, 300, 350, 400, 450, 500, 550, 600, 700, 800, 900,1000, 1100, 1200, 1300, 1400, and/or 1500 psig, (2) less than or equalto 3000, 2000, 1900, 1800, 1700, 1600, 1500, 1400, 1300, 1200, 1100,1000, 900, 800, 700, 600, 500, 400, and/or 350, and/or (3) between −14.5and 3000 psig, between 0 and 1500 psig, and/or within any range boundedby any combination of the upper and lower numbers and/or any nestedrange within such ranges.

According to various embodiments, the outlet temperature of the workingfluid when the working fluid is expelled from the compression chamberexceeds the inlet temperature of the working fluid when the workingfluid enters the compression chamber by (a) less than 700, 650, 600,550, 500, 450, 400, 375 350, 325, 300, 275, 250, 225, 200, 175, 150,140, 130, 120, 110, 100, 90, 80, 70, 60, 50, 40, 30, and/or 20 degreesC., (b) at least −10, 0, 10, and/or 20 degrees C., and/or (c) anycombination of ranges between any two of these upper and lower numbers,including any range within such ranges.

According to various embodiments, the outlet temperature of the workingfluid is (a) less than 700, 650, 600, 550, 500, 450, 400, 375, 350, 325,300, 275, 250, 225, 200, 175, 150, 140, 130, 120, 110, 100, 90, 80, 70,60, 50, 40, 30, and/or 20 degrees C., (b) at least −10, 0, 10, 20, 30,40, and/or 50 degrees C., and/or (c) any combination of ranges betweenany two of these upper and lower numbers, including any range withinsuch ranges.

The outlet temperature and/or temperature increase may be a function ofthe working fluid. For example, the outlet temperature and temperatureincrease may be lower for some working fluids (e.g., methane) than forother working fluids (e.g., air).

According to various embodiments, the temperature increase is correlatedto the pressure ratio. According to various embodiments, the temperatureincrease is less than 200 degrees C. for a pressure ratio of 20:1 orless (or between 15:1 and 20:1), and the temperature increase is lessthan 300 degrees C. for a pressure ratio of between 20:1 and 30:1.

According to various embodiments, the pressure ratio is between 3:1 and15:1 for a working fluid with an inlet liquid volume fraction of over5%, and the pressure ratio is between 15:1 and 40:1 for a working fluidwith an inlet liquid volume fraction of between 1 and 20%. According tovarious embodiments, the pressure ratio is above 15:1 while the outletpressure is above 250 psig, while the temperature increase is less than200 degrees C. According to various embodiments, the pressure ratio isabove 25:1 while the outlet pressure is above 250 psig and thetemperature increase is less than 300 degrees C. According to variousembodiments, the pressure ratio is above 15:1 while the outlet pressureis above 250 psig and the compressor speed is over 450 rpm.

According to various embodiments, any combination of the differentranges of different parameters discussed herein (e.g., pressure ratio,inlet temperature, outlet temperature, temperature change, inletpressure, outlet pressure, pressure change, compressor speed, coolantinjection rate, etc.) may be combined according to various embodimentsof the invention. According to one or more embodiments, the pressureratio is anywhere between 3:1 and 200:1 while the operating compressorspeed is anywhere between 350 and 3000 rpm while the outlet pressure isbetween 200 and 6000 psig while the inlet pressure is between 0 and 3000psig while the outlet temperature is between −10 and 650 degrees C.while the outlet temperature exceeds the inlet temperature by between 0and 650 degrees C. while the liquid volume fraction of the working fluidat the compressor inlet is between 1% and 50%.

According to one or more embodiments, air is compressed from ambientpressure (14.7 psia) to 385 psia, a pressure ratio of 26:1, at speeds of700 rpm with outlet temperatures remaining below 100 degrees C. Similarcompression in an adiabatic environment would reach temperatures ofnearly 480 degrees C.

The operating speed of the illustrated compressor is stated in terms ofrpm because the illustrated compressor is a rotary compressor. However,other types of compressors may be used in alternative embodiments of theinvention. As those familiar in the art appreciate, the RPM term alsoapplies to other types of compressors, including piston compressorswhose strokes are linked to RPM via their crankshaft.

Numerous cooling liquids may be used. For example, water, triethyleneglycol, and various types of oils and other hydrocarbons may be used.Ethylene glycol, propylene glycol, methanol or other alcohols in casephase change characteristics are desired may be used. Refrigerants suchas ammonia and others may also be used. Further, various additives maybe combined with the cooling liquid to achieve desired characteristics.Along with the heat transfer and heat absorption properties of theliquid helping to cool the compression process, vaporization of theliquid may also be utilized in some embodiments of the design to takeadvantage of the large cooling effect due to phase change.

The effect of liquid coalescence is also addressed in the preferredembodiments. Liquid accumulation can provide resistance against thecompressing mechanism, eventually resulting in hydrolock in which allmotion of the compressor is stopped, causing potentially irreparableharm. As is shown in the embodiments of FIGS. 8 and 17, the inlet 420and outlet 430 are located at the bottom of the rotor casing 400 onopposite sides of the gate 600, thus providing an efficient location forboth intake of fluid to be compressed and exhausting of compressed fluidand the injected liquid. A valve is not necessary at the inlet 420. Theinclusion of a dwell seal allows the inlet 420 to be an open port,simplifying the system and reducing inefficiencies associated with inletvalves. However, if desirable, an inlet valve could also beincorporated. Additional features may be added at the inlet to induceturbulence to provide enhanced thermal transfer and other benefits.Hardened materials may be used at the inlet and other locations of thecompressor to protect against cavitation when liquid/gas mixtures enterinto choke and other cavitation-inducing conditions.

Alternative embodiments may include an inlet located at positions otherthan shown in the figures. Additionally, multiple inlets may be locatedalong the periphery of the cylinder. These could be utilized inisolation or combination to accommodate inlet streams of varyingpressures and flow rates. The inlet ports can also be enlarged or moved,either automatically or manually, to vary the displacement of thecompressor.

In these embodiments, multi-phase compression is utilized, thus theoutlet system allows for the passage of both gas and liquid. Placementof outlet 430 near the bottom of the rotor casing 400 provides for adrain for the liquid. This minimizes the risk of hydrolock found inother liquid injection compressors. A small clearance volume allows anyliquids that remain within the chamber to be accommodated. Gravityassists in collecting and eliminating the excess liquid, preventingliquid accumulation over subsequent cycles. Additionally, the sweepingmotion of the rotor helps to ensure that most liquid is removed from thecompressor during each compression cycle by guiding the liquid towardthe outlet(s) and out of the compression chamber.

Compressed gas and liquid can be separated downstream from thecompressor. As discussed below, liquid coolant can then be cooled andrecirculated through the compressor.

Various of these features enable compressors according to variousembodiments to effectively compress multi-phase fluids (e.g., a fluidthat includes gas and liquid components (sometimes referred to as “wetgas”)) without pre-compression separation of the gas and liquid phasecomponents of the working fluid. As used herein, multi-phase fluids haveliquid volume fractions at the compressor inlet port of (a) at least0.5, 1, 2, 3, 4, 5, 6, 7, 8, 9, 10, 15, 20, 25, 30, 35, 40, 50, 60, 70,75, 80, 85, 90, 91, 92, 93, 94, 95, 96, 97, 98, 99, and/or 99.5%, (b)less than or equal to 99.5, 99, 98, 97, 96, 95, 94, 93, 92, 91, 90, 85,80, 75, 70, 60, 50, 40, 35, 30, 25, 20, 15, 10, 9, 8, 7, 6, 5, 4, 3, 2,1, and/or 0.5%, (c) between 0.5 and 99.5%, and/or (d) within any rangebounded by these upper and lower values.

Outlet valves allow gas and liquid (i.e., from the wet gas and/or liquidcoolant) to flow out of the compressor once the desired pressure withinthe compression chamber is reached. The outlet valves may increase ormaximize the effective orifice area. Due to the presence of liquid inthe working fluid, valves that minimize or eliminate changes indirection for the outflowing working fluid are desirable, but notrequired. This prevents the hammering effect of liquids as they changedirection. Additionally, it is desirable to minimize clearance volume.Unused valve openings may be plugged in some applications to furtherminimize clearance volume. According to various embodiments, thesefeatures improve the wet gas capabilities of the compressor as well asthe compressor's ability to utilize in-chamber liquid coolant.

Reed valves may be desirable as outlet valves. As one of ordinary skillin the art would appreciate, other types of valves known or as yetunknown may be utilized. Hoerbiger type R, CO, and Reed valves may beacceptable. Additionally, CT, HDS, CE, CM or Poppet valves may beconsidered. Other embodiments may use valves in other locations in thecasing that allow gas to exit once the gas has reached a given pressure.In such embodiments, various styles of valves may be used. Passive ordirectly-actuated valves may be used and valve controllers may also beimplemented.

In the presently preferred embodiments, the outlet valves are locatednear the bottom of the casing and serve to allow exhausting of liquidand compressed gas from the high pressure portion. In other embodiments,it may be useful to provide additional outlet valves located alongperiphery of main casing in locations other than near the bottom. Someembodiments may also benefit from outlets placed on the endplates. Instill other embodiments, it may be desirable to separate the outletvalves into two types of valves—one predominately for high pressuredgas, the other for liquid drainage. In these embodiments, the two ormore types of valves may be located near each other, or in differentlocations.

The coolant liquid can be removed from the gas stream, cooled, andrecirculated back into the compressor in a closed loop system. Byplacing the injector nozzles at locations in the compression chamberthat do not see the full pressure of the system, the recirculationsystem may omit an additional pump (and subsequent efficiency loss) todeliver the atomized droplets. However, according to alternativeembodiments, a pump is utilized to recirculate the liquid back into thecompression chamber via the injector nozzles. Moreover, the injectornozzles may be disposed at locations in the compression chamber that seethe full pressure of the system without deviating from the scope ofvarious embodiments of the present invention.

According to various embodiments, some compressed working fluid/gas(e.g., natural gas) that has been compressed by the compressor isrecirculated back into the compression chamber via the injector nozzlesalong with coolant to better atomize the coolant (e.g., similar oridentical to how snow-making equipment combines a liquid water streamwith a compressed gas stream to achieve increase atomization of thewater).

One or more embodiments simplify heat recovery because most or all ofthe heat load is in the cooling liquid. According to variousembodiments, heat is not removed from the compressed gas downstream ofthe compressor. The cooling liquid may cooled via an active coolingprocess (e.g., refrigeration and heat exchangers) downstream from thecompressor. However, according to various embodiments, heat mayadditionally be recovered from the compressed gas (e.g., via heatexchangers) without deviating from the scope of various embodiments ofthe present invention.

As shown in FIGS. 8 and 17, the sealing portion 510 of the rotoreffectively precludes fluid communication between the outlet and inletports by way of the creation of a dwell seal. The interface between therotor 500 and gate 600 further precludes fluid communication between theoutlet and inlet ports through use of a non-contacting seal or tip seal620. In this way, the compressor is able to prevent any return andventing of fluid even when running at low speeds. Existing rotarycompressors, when running at low speeds, have a leakage path from theoutlet to the inlet and thus depend on the speed of rotation to minimizeventing/leakage losses through this flow path.

The high pressure working fluid exerts a large horizontal force on thegate 600. Despite the rigidity of the gate struts 210, this force willcause the gate 600 to bend and press against the inlet side of the gatecasing 152. Specialized coatings that are very hard and have lowcoefficients of friction can coat both surfaces to minimize friction andwear from the sliding of the gate 600 against the gate casing 152. Afluid bearing can also be utilized. Alternatively, pegs (not shown) canextend from the side of the gate 600 into gate casing 150 to helpsupport the gate 600 against this horizontal force. Material may also beremoved from the non-pressure side of gate 600 in a non-symmetricalmanner to allow more space for the gate 600 to bend before interferingwith the gate casing 150.

The large horizontal forces encountered by the gate may also requireadditional considerations to reduce sliding friction of the gate'sreciprocating motion. Various types of lubricants, such as greases oroils may be used. These lubricants may further be pressurized to helpresist the force pressing the gate against the gate casing. Componentsmay also provide a passive source of lubrication for sliding parts vialubricant-impregnated or self-lubricating materials. In the absence of,or in conjunction with, lubrication, replaceable wear elements may beused on sliding parts to ensure reliable operation contingent onadherence to maintenance schedules. These wear elements may also be usedto precisely position the gate within the gate casing. As one ofordinary skill in the art would appreciate, replaceable wear elementsmay also be utilized on various other wear surfaces within thecompressor.

The compressor structure may be comprised of materials such as aluminum,carbon steel, stainless steel, titanium, tungsten, or brass. Materialsmay be chosen based on corrosion resistance, strength, density, andcost. Seals may be comprised of polymers, such as PTFE, HDPE, PEEK™,acetal copolymer, etc., graphite, cast iron, carbon steel, stainlesssteel, or ceramics. Other materials known or unknown may be utilized.Coatings may also be used to enhance material properties.

As one of ordinary skill in the art can appreciate, various techniquesmay be utilized to manufacture and assemble embodiments of the inventionthat may affect specific features of the design. For example, the maincasing 110 may be manufactured using a casting process. In thisscenario, the nozzle housings 132, gate casing 150, or other componentsmay be formed in singularity with the main casing 110. Similarly, therotor 500 and drive shaft 140 may be built as a single piece, either dueto strength requirements or chosen manufacturing technique.

Further benefits may be achieved by utilizing elements exterior to thecompressor envelope. A flywheel may be added to the drive shaft 140 tosmooth the torque curve encountered during the rotation. A flywheel orother exterior shaft attachment may also be used to help achievebalanced rotation. Applications requiring multiple compressors maycombine multiple compressors on a single drive shaft with rotors mountedout of phase to also achieve a smoothened torque curve. A bell housingor other shaft coupling may be used to attach the drive shaft to adriving force such as engine or electric motor to minimize effects ofmisalignment and increase torque transfer efficiency. Accessorycomponents such as pumps or generators may be driven by the drive shaftusing belts, direct couplings, gears, or other transmission mechanisms.Timing gears or belts may further be utilized to synchronize accessorycomponents where appropriate.

After exiting the valves the mix of liquid and gases may be separatedthrough any of the following methods or a combination thereof: 1.Interception through the use of a mesh, vanes, intertwined fibers; 2.Inertial impaction against a surface; 3. Coalescence against otherlarger injected droplets; 4. Passing through a liquid curtain; 5.Bubbling through a liquid reservoir; 6. Brownian motion to aid incoalescence; 7. Change in direction; 8. Centrifugal motion forcoalescence into walls and other structures; 9. Inertia change by rapiddeceleration; and 10. Dehydration through the use of adsorbents orabsorbents.

At the outlet of the compressor, a pulsation chamber may consist ofcylindrical bottles or other cavities and elements, may be combined withany of the aforementioned separation methods to achieve pulsationdampening and attenuation as well as primary or final liquidcoalescence. Other methods of separating the liquid and gases may beused as well.

FIGS. 39-44 illustrate a compressor 1000 according to an alternativeembodiment. The compressor 1000 is generally similar to theabove-discussed compressors. Accordingly, a redundant description ofsimilar or identical components is omitted. The compressor 1000 includesa main casing 1010 that defines a compression chamber 1020, a driveshaft 1030, a rotor 1040, cams 1050, cam followers 1060, a gate support1070 (e.g., cam follower supports, cam struts, gate support arm, gatestrut, etc.) connected to the cam followers 1060, a gate support guide1075 mounted to the casing 1010 (or integrally formed with the casing1010) and connected to the gate support 1070 to permit reciprocal linearmovement of the gate support 1070, springs 1080 that bias the gatesupport 1070 toward the cams 1050, a gate housing 1100 that is partiallyformed by and/or mounted to the main casing 1010 and/or the gate supportguide 1075, a gate 1110 slidingly supported by the gate housing 1100, aninlet manifold 1140 fluidly connected to an inlet 1150 into thecompression chamber 1020, a discharge/outlet manifold 1160 fluidlyconnected to a discharge outlet 1170 that leads from the compressionchamber 1020, a discharge outlet valve 1180 disposed in the dischargeoutlet 1170, coolant injectors 1190, a hydrostatic bearing arrangement1300 (see FIGS. 48-51) between the casing 1010 and gate 1110, and amechanical/hydraulic seal 1500 that seals the compression chamber 1020from the ambient environment around the drive shaft 1030.

In the illustrated embodiment, the coolant injectors 1190 direct coolantdirectly into the compression chamber 1020. However, according to one ormore alternative embodiments, coolant injector(s) 1190 may additionallyand/or alternatively inject coolant into the working fluid in the inletmanifold 1140 before the working fluid or coolant reach the compressionchamber. Such an alternative may reduce manufacturing costs and/orreduce the amount of power required to inject the coolant.

As shown in FIGS. 41, 43, and 44, the discharge outlet valve 1180directs compressed fluid through the discharge outlet 1170 whilediscouraging backflow of compressed fluid back into the compressionchamber 1020. As shown in FIG. 41, the valve 1180 is separately formedfrom the main casing 1010 and is fitted into the discharge outlet 1170.However, according to various alternative embodiments, the valve 1180 orparts thereof may be integrally formed with the casing 1010.

As shown in FIGS. 45-46, the discharge manifold 1160 includes aplurality of vanes 1160 a. A cross-section of a passageway within themanifold 1160 from the discharge outlet 1170 (i.e., entrance into themanifold 1160) to a circular discharge manifold outlet 1160 b (i.e., adownstream exit of the manifold 1160) transitions from anaxially-elongated cross-section at the discharge outlet 1170 (e.g.,elongated along the length of the gate 1110 in a direction parallel tothe rotational axis of the drive shaft 1030) to the circular dischargemanifold outlet 1160 b. According to various embodiments, thecross-sectional area remains relatively constant throughout thisdischarge flow path. The vanes 1160 a are oriented generallyperpendicular to the desired flow path of the compressed fluid from thecompression chamber 1020 to a discharge manifold outlet 1160 b of thedischarge manifold 1160. The vanes 1160 a are oriented to promote agenerally laminar flow of the compressed fluid as the cross-sectionalshape of the flow path changes. According to various embodiments, thevanes 1160 a reduce turbulence, increase the efficiency of thecompressor 1000, and/or reduce wear as the compressed fluid (e.g.,multiphase liquid/gas fluid) flows though the outlet 1170 and manifold1160.

The vanes 1160 a and valve 1180 extend completely across the flow pathof compressed fluid (e.g., into the page as shown in FIG. 45, up anddown as shown in FIG. 47, from an upper left toward a lower right asshown in FIG. 43). The vanes 1160 a and valve 1180 thereforestructurally support circumferentially-spaced portions 1010 a, 1010 b(see FIG. 43) of the casing 1010 on either side of the axially-elongateddischarge outlet 1170. The vanes 1160 a and valve 1180 may thereforehelp the casing 1010 to resist deformation (e.g., that might beencouraged by reaction forces generated between the gate 1110 and casing1010 during use of the compressor 1000).

As shown in FIG. 48, a plurality of vanes/ribs 1155 are disposed withinand extend across the inlet 1150 along the circumferential direction ofthe compression chamber 1020 (from lower left to upper right as shown inFIG. 48). These ribs 1155 strengthen the casing 1010 in the area of theinlet 1150, and help to prevent deflection of the casing 1010 around thegate 1110. According to various embodiments, the inlet 1150 is axiallydivided into a plurality of discrete inlets 1150 (e.g., holes spacedalong the axial direction of the compressor 1000), such that thevanes/ribs 1155 are defined by portions of the casing 1010 between suchinlet holes.

As illustrated in FIGS. 48-51, the compressor 1000 includes ahydrostatic bearing arrangement 1300 that allows the gate 1110 toreciprocate up and down relative to the gate housing 1100 whilemaintaining close contact with the rotor 1040. The hydrostatic bearingarrangement 1300 reduces friction between the gate 1110 and the gatehousing 1100.

As shown in FIGS. 43, 48 and 50, the gate 1110 separates an inlet side1020 a of the compression chamber 1020 from an outlet side 1020 b of thecompression chamber 1020. Pressure in the inlet side 1020 a staysrelatively close to the pressure of fluid entering the compressionchamber 1020 via the inlet 1150. Pressure in the outlet side 1020 b ofthe compression chamber 1020 increases during each compressionstroke/revolution and reaches the output pressure of compressed fluidbeing output through the discharge outlet 1170. As shown in FIG. 50,this causes a higher pressure on the outlet side 1020 b of the gate 1110than on the inlet side 1020 a, which pushes the gate toward the inletside 1020 a. As shown in FIG. 50, this differential pressure creates acantilever force on the gate 1110 and because the compression chamber1020 pressure increases until discharge every cycle the cantilever forceis constantly cycling. The hydrostatic bearing arrangement 1300accommodates this cycling cantilever force and equalizes thecantilever/bending moment on the gate 1110.

As shown in FIGS. 48-51, the hydrostatic bearing arrangement 1300comprises: upper hydrostatic bearings 1310 on the inlet side 1020 a ofthe gate 1110, lower hydrostatic bearings 1320 on the inlet side 1020 aof the gate 1110, upper hydrostatic bearings 1330 on thecompression/outlet side 1020 b of the gate 1110, and lower hydrostaticbearings 1340 on the compression/outlet side 1020 b of the gate 1110.

As shown in FIG. 49, three of each bearing 1310, 1320, 1330, 1340 arespaced apart along the axial/longitudinal direction of the compressor1000 (i.e., into the page as shown in FIG. 50), such that there arethree columns of bearings 1310, 1320, 1330, 1340 (or six columns if bothsides 1020 a, 1020 b are considered separate). According to variousnon-limiting embodiments, the use of multiple columns of bearings 1310,1320, 1330, 1340 may reduce the length the hydraulic fluid has tolaterally travel. This may keep hydraulic fluid more evenly distributedover all surfaces of the bearing pad. Increasing the number of bearingsmay also isolate problems (e.g., debris, deflection of bearing surfaces,wear of bearing pad surfaces, a clog in the oil system, etc.) to asingle bearing 1310, 1320, 1330, 1340 leaving other bearings 1310, 1320,1330, 1340 still working properly. However, greater or fewer columns ofbearings 1310, 1320, 1330, 1340 could be used without deviating fromvarious embodiments (e.g., by combining the different bearings 1310 intoa single longitudinally longer bearing). According to one or moreembodiments, four columns of bearings are provided on each side of thegate.

According to various embodiments, the use of multiple columns ofbearings 1310, 1320, 1330, 1340 may facilitate fine tuning of theresistors 1410 of one column (or bearings within one column) relative toother column(s) to accommodate for varying conditions along the lengthof the gate 1110. For example, if the hydrostatic pressure causes thesleeve 1360 to bow out in the middle, the middle column of bearings1310, 1320, 1330, 1340 can be tuned down to decrease flow to thoselarger gaps and increase flow to the end columns where the gaps aretighter and contact between the gate and sleeve would first be made.

As shown in FIGS. 48-50, the hydrostatic bearing arrangement 1300 isformed in a hydrostatic bearing insert/sleeve 1360 that mates with thecasing 1010. Shims or other suitable mechanisms may be used to ensure asecure, low-tolerance fit and positioning of the sleeve 1360. The sleeve1360 is removable from the casing 1010 to facilitate replacement ofand/or maintenance on the sleeve 1360. However, according to alternativeembodiments, the insert 1360 may be integrally formed with the casing1010.

As shown in FIG. 51, each bearing 1310, 1320, 1330, 1340 comprises aninlet port 1310 a, 1320 a, 1330 a, 1340 a that opens into a pocketgroove 1310 b, 1320 b, 1330 b, 1340 b on a side of the insert 1360 thatmates with the gate 1110. Each groove 1310 b, 1320 b, 1330 b, 1340 b issurrounded by a land/bearing pad 1310 c, 1320 c, 1330 c, 1340 c thatclosely mates with the gate 1110. The pad 1310 c, 1320 c, 1330 c, 1340 cis surrounded by a drain 1370, which may be common to all of thebearings 1310, 1320, 1330, 1340.

As shown in FIG. 51, a hydraulic pump 1380 pumps hydraulic fluid (e.g.,oil) from a reservoir 1390 through hydraulic passageways 1400 torespective resistor flow valves 1410 for each of the bearings 1310,1320, 1330, 1340. The passageways 1400 then lead sequentially torespective inlet ports 1310 a, 1320 a, 1330 a, 1340 a, grooves 1310 b,1320 b, 1330 b, 1340 b, lands/bearing pads 1310 c, 1320 c, 1330 c, 1340c, the drain 1370, and back into the reservoir 1390.

As already known, hydrostatic bearings work by using two flow resistors.In this embodiment, the first flow resistor is a flow resistor valve1410 inline prior to the bearing 1310, 1320, 1330, 1340, which is heldconstant during operation. The bearing pad 1310 c, 1320 c, 1330 c, 1340c itself is the second flow resistor. The resistance of the bearing pad1310 c, 1320 c, 1330 c, 1340 c changes and is dependent on the gapbetween the gate 1110 and the bearing pad itself 1310 c, 1320 c, 1330 c,1340 c. If this gap decreases the pressure in the bearing pad 1310 c,1320 c, 1330 c, 1340 c and the pocket grooves 1310 b, 1320 b, 1330 b,1340 b will go up and similarly if the gap increases the pressure in thepad 1310 c, 1320 c, 1330 c, 1340 c and the pocket grooves 1310 b, 1320b, 1330 b, 1340 b will go down. The gap will change due to loads createdby the cantilever pressure force on the gate 1110.

According to various embodiments, the flow resistor valve 1410 can bereplaced by a set flow resistor or an annulus in the respectivepassageway 1400 that behaves similarly to the bearing pad resistor. Anannulus can be designed into the bearing pad 1310 c, 1320 c, 1330 c,1340 c that allows flow to pass through it with a resistance that isdependent on the gap. Typically the annulus is placed on the oppositesurface of the bearing pad to which it is hydraulically connected. To beclear, lubricant would flow through the annulus on one side of thebearing and then flow to its respective bearing pad on the oppositeside. Thus, according to various embodiments, the bearings 1310, 1320,1330, 1340 comprise self-compensating bearings with flow resistors builtinto the opposing bearings. For example, the flow resistor valve 1400for the bearing 1310 may be built into the opposite bearing 1330 so thatflow to the bearing 1310 is reduced when the bearing 1330 gap isreduced. This may prevent excess hydraulic fluid flow through bearings1310, 1320, 1330, 1340 with large gaps (because the gap on the opposingbearing is small) or permit larger flow rates to bearings 1310, 1320,1330, 1340 that have higher loads. Bearings 1320, 1340 oppose each otherand can work in the same manner. This type of self-compensatinghydrostatic bearing is described in U.S. Pat. No. 7,287,906, the entirecontents of which are incorporated herein by reference.

As shown in FIG. 50, according to various embodiments, the use of upperbearings 1310, 1330 that are discrete from lower bearings 1320, 1340enables the bearing arrangement 1300 to adapt to the cantilever/bendingmoments being exerted on the gate 1110 by the pressurized fluid in thecompression chamber 1020, 1020 b and the rotor 1040. The magnitude ofthe forces being exerted on the gate 1110 by the inlet and outlet sides1020 a, 1020 b of the compression chamber 1020 and the bearings 1310,1320, 1330, 1340 is represented by the size of the arrows. As shown inFIG. 50, when the outlet side 1020 b force is high relative to the inletside 1020 a, the moment is balanced by a high force from the upperfar-side bearing 1310 and lower near-side bearing 1340, where the gapsare the smallest. Conversely, the bearing gaps are larger between thegate 1110 and bearings 1320, 1330, such that the force applied by thesebearings 1320, 1330 is lower. According to various alternativeembodiments, additional upper, lower, and/or intermediate hydrostaticbearings may be added to more specifically account for the bendingmoment being exerted on the gate 1110. However, according to alternativeembodiments, the upper and lower hydrostatic bearings (e.g., bearings1330, 1340; bearings 1310, 1320) may be combined without deviating fromthe scope of various embodiments.

As used herein, the directional terms “upper” and “lower” with respectto bearings 1310, 1330, 1320, 1340 are defined along the direction ofreciprocating movement of the gate 1110, and not necessarily along agravitational up/down direction (though gravitational up/down alignswith the gate 1110's up/down reciprocating direction according tovarious embodiments).

According to various embodiments, the hydrostatic bearing arrangement1300 creates a fluid film gap between the gate 1110 and casing 1010 onthe inlet side 1020 a of the compression chamber 1020, which may prolongthe useful life of the gate 1110 and/or casing 1010 by reducing oreliminating wearing contact between the gate 1110 and casing 1010,and/or reduce the forces required to move the gate 1110 along itsreciprocating path.

According to various alternative embodiments, the hydrostatic bearing isused on a rotary vane compressor in which the vanes rotate with andreciprocate relative to the rotor instead of the casing. In suchembodiments, a hydrostatic bearing such as the bearing 1300 is disposedbetween the rotor and gate, rather than between the casing and gate.

As shown in FIG. 50, the gate 1110 includes a seal 1430 that mounts to agroove 1440 a in the main body 1440 of the gate 1110. As shown in FIG.50, the seal 1430 and groove 1440 a have complimentary “+” shapedprofiles that help to retain the seal 1430 in the groove 1440 a duringoperation of the compressor 1000. According to various alternatives, thegroove 1440 a and seal 1430 may have any other suitable complimentaryprofile that discourages separation of the seal 1430 from the gate body1440 (e.g., a profile with a narrow top opening and a larger (e.g.,bulbous) middle cross-section, a triangular profile with a point towardthe top, etc.).

As shown in FIG. 50, according to various embodiments, the gate body1440 and/or the sleeve 1360 may be formed from hard materials thatresist wear (e.g., materials such as 440C steel, 17-4 steel, D2 toolsteel, or Inconel, among others, with HRC over 35, 40, 45, 50, 55, 60,65, etc.) or are coated with wear-resistant coatings or otherwisetreated to increase hardness (e.g., nitrided steel, steel with a hardceramic coating, steel with surface heat treatments that increasesurface hardness, etc.) so as to resist wear when and if the sleeve 1360and gate body 1440 rub against each other. Additionally and/oralternatively, one of the sleeve 1360 and gate body 1440 may have a hardsurface (e.g., steel) while the other of the sleeve 1360 and gate body1440 is relatively softer (e.g., formed of bronze of brass) so as to besacrificially worn during operation, and eventually replaced. Accordingto one or more embodiments, the sleeve 1360 comprises a hard-surfacedmaterial such as steel, while the gate body 1440 comprises a softmaterial such as bronze. According to one or more alternativeembodiments, the sleeve 1360 comprises a soft material such as bronze,while the gate body 1440 comprises a hard material such as steel.

According to various embodiments, the surface of the gate 1110 and/orsleeve 1360 (or a coating thereon) is matted or otherwise constructed soas to create turbulence within the oil flow, thereby increasing theshear force of the oil as it forces its way through the gaps andincreases the hydrostatic bearing pressure.

According to alternative embodiments, the hydrostatic bearingarrangement 1300 is replaced with a hydrodynamic bearing arrangement,which provides hydraulic liquid (e.g., oil) to an interface between thegate body 1440 and sleeve 1360. The hydrodynamic bearing relies onrelative movement between the gate body 1440 and sleeve 1360 to causethe hydraulic fluid to pressurize and/or lubricate the intersection.

As shown in FIG. 40, a mechanical seal 1500 on each axial end of thecompressor 1000 hermitically seals the compression chamber 1020 of thecompressor 1000 relative to the environment outside of the compressionchamber 1020 around the driveshaft 1030.

Each of the two mechanical seals 1500 includes face seals 1510, 1520, aradial shaft seal 1550, a vent 1560, and hydraulic packing 1590. Asshown in FIGS. 40, 52, and 54, the inner and outer face seals 1510, 1520seal an axial end of the rotor 1040 relative to the axial face of thecasing 1010 that defines the compression chamber 1020. As shown in FIG.52, the seals 1510, 1520 are mounted within circumferential (butnon-circular in the case of seal 1520) face grooves 1040 b in the rotor1040 to permit axial movement (i.e., left/right movement as shown inFIG. 40), and springs 1530, 1540 (e.g., Belleville washers, an O-ringwith elastic properties, a series of compression springs arranged aroundthe perimeter of the seals 1501, 1520) bias the seals 1510, 1520 axiallyagainst the axial face of the casing 1010 that defines the compressionchamber 1020. The inner face seal 1510 is circular and concentric with arotational axis of the drive shaft 1030. As shown in FIG. 41, the outerface seal 1520 follows the non-circular perimeter of the rotor 1040, androtates with the rotor 1040 about the axis of the drive shaft 1030.According to various embodiments, outer sealing portions of the faceseals 1510, 1520 comprise low-friction material (e.g., graphite) that isbonded to a stronger backing (e.g., steel).

According to various embodiments, the seals 1510, 1520 are retained intheir grooves 1040 b even when the wear surface of the seals 1510, 1520(e.g., the graphite portion of the seals 1510, 1520) is worn through.For example, as shown in FIGS. 67 and 68, the seals 1510, 1520 may beretained by locking washers 1541 (e.g., multiple washers per seal 1510,1520) that are connected (e.g., via bolts 1542 or other fasteners) torecesses 1040 c in the end faces of the rotor 1040 and extend intoshouldered grooves 1510 a, 1520 a in the seals 1510, 1520 to prevent theseals 1510, 1520 from separating from mating seal grooves 1040 b, whilepermitting the seals 1510, 1520 to move axially within the grooves 1040b to keep the seals 1510, 1520 proximate to the mating face of thecompression chamber (e.g., the face of wear plate 1545 (see FIG. 52).

As shown in FIG. 52, an end cap wear plate 1545 on each axial end of thecompression chamber 1020 removably mounts to a remainder of the casing1010 (e.g., via bolts) and abuts the seals 1510, 1520. The plate 1545may be replaced when wearing contact between the seals 1510, 1520 andplate 1545 has worn the plate 1545 sufficiently to warrant replacement.

As shown in FIG. 54, the radial shaft seal 1550 extends radially betweenthe drive shaft 1030 and an end cap of the casing 1010. As shown inFIGS. 54 and 40, the vent 1560 is disposed axially outwardly from theradial shaft seals 1550. As shown in FIG. 54, a fluid passageway 1570fluidly connects the vent 1560 to the inlet 1150 of the compressor 1000.As shown in FIG. 54, the hydraulic packing 1590 comprises facing radialseals 1600, 1610 with a hydraulic fluid passage 1620 therebetween. Thehydraulic pump 1380 (or any other suitable source of hydraulic fluid)provides pressurized hydraulic fluid to the hydraulic packing 1590 via aport/passageway 1630 that leads into the space between the seals 1600,1610. As shown in FIG. 54, rotational bearings 1650 support the driveshaft 1030 relative to the casing 1010 to permit the drive shaft 1030 torotate relative to the casing 1010.

The operation of the mechanical seal 1500 is described with reference toFIGS. 52 and 54. For the working fluid (e.g., natural gas beingcompressed) to leak out of the compression chamber 1020, the fluid mayleak sequentially through the seals 1520, 1510, 1550. If the workingfluid leaks past all three seals 1520, 1510, 1550, the fluid reaches thevent 1560 which returns the fluid back to the compressor inlet 1150 viathe passageway/port 1570, which is maintained at the pressure of theinlet 1150 via its fluid communication with the inlet 1150. Thehydraulic packing 1590 on the outer axial side of the vent 1560 ispressurized via hydraulic fluid to a pressure higher than the inlet 1150pressure, which discourages or prevents the working fluid from furtherleaking past the hydraulic packing 1590. Leaked working fluid leaksthrough the passageway/port 1570 back to the intake 1150, rather thanpast the hydraulic packing 1590 because the inlet 1150 is at asignificantly lower pressure than the hydraulic packing 1590. Thus,leakage of the working fluid past the hydraulic packing 1590 is reducedor preferably eliminated. Pressure in the bearing cavity for thebearings 1650 is maintained at ambient atmospheric pressure.

According to various embodiments, the mechanical seal 1500 provides anaxially-compact seal that results in lower moment loads on thecompressor's bearings.

As shown in FIG. 52, in the compressor 1000, the drive shaft 1030 ismounted to each axial end of the casing 1010 via a combination ofseparate rotational bearings 1650 and thrust bearings 1660. However, asshown in FIG. 53, the separate rotational and thrust bearings 1650, 1660may be replaced by a consolidated bearing 1670 that serves both thrustbearing and rotational bearing functions without deviating from thescope of various embodiments. To facilitate removal of the bearing 1670from the drive shaft, a lubrication passageway may extend through thedrive shaft and open into the interface between the drive shaft and thebearing 1670. According to various alternative embodiments, the bearings1650, 1660 may be replaced with any other type of rotational couplingbetween the drive shaft 1030 and casing 1010 without deviating from thescope of various embodiments (e.g., other types of bearings, bushings,etc.).

Although the seal 1500 is described as including various structures inthe illustrated embodiment, the seal 1500 may include greater or fewerstructures without deviating from the scope of the present invention.For example, one or more of the seals 1510, 1520, 1550 may be omittedwithout deviating from the scope of the present invention.

FIG. 69 illustrates a compressor 5150 that is generally similar to thecompressor 1000, except that the compressor 5150 uses an alternativeembodiment of a mechanical seal 5200 in place of the mechanical seal1500. The mechanical seal 5200 is generally similar to the seal 1500, soa redundant explanation of similar or identical components is omitted.In contrast with the axially spaced arrangement of various components ofthe mechanical seal 1500 (e.g., the radial seal 1550, vent 1560, radialseals 1600, 1610, and pressurized hydraulic fluid passageway 1620),various components of the mechanical seal 5200 are radially spaced fromeach other, which may provide a more axially-compact seal. As shown inFIG. 69, the compressor 5150 includes a casing 5210 that is generallyidentical to the casing 1010, except that the casing 5210 is shapedslightly differently so as to accommodate the differently shapedmechanical seal 5200.

As shown in FIG. 69, the seal 5200 includes an annular collar 5220 thatis rigidly and sealingly connected to or integrally formed with thedrive shaft 1030 so as to rotate with the drive shaft 1030 relative tothe casing 5210. According to various embodiments, the collar 5220 mayconnect to the drive shaft 1030 in a variety of alternative ways (e.g.,heat-shrunk onto the shaft 1030, glued or otherwise fastened onto theshaft 1030, welded onto the shaft 1030, press-fit onto the shaft 1030,etc.). According to various embodiments, o-rings 5230 are disposedbetween the collar 5220 and shaft 1030 to prevent leaks therebetween.Inner annular seal grooves 5220 a,b and outer annular seal grooves 5220c,d are disposed on the axial faces of the collar 5220 that face towardand away from the rotor 1040. Face seals 5240, 5250, 5260, 5270 aredisposed in the grooves 5220 a,b,c,d and spring biased away from thecollar 5220 toward a mating axial face surface 5210 a, 5210 b of thecasing 5210. A vent 5290 is disposed between the collar 5220 and casing5210 radially outwardly from the collar 5220. The vent 5290 fluidlyconnects to an inlet into the compressor 5150 via a passageway 5300 inthe casing 5210. A hydraulic fluid passageway 5310 connects a source ofpressurized hydraulic fluid (or other fluid) (e.g., the pump 1380) to aspace 5330 disposed between the seals 5250, 5270, face 5210 b, andcollar 5220 so as to keep this space 5330 pressurized with hydraulicfluid.

The operation of the mechanical seal 5200 is described with reference toFIG. 69. If working fluid leaks from the compression chamber 1020sequentially past the face seal 1520, face seal 1510, face seal 5240,and face seal 5260, the leaked working fluid will leak into the vent5290, which will direct the leaked working fluid back to the inlet ofthe compressor 5150 via the passageway 5300. As with the seal 1500, thehydraulic packing formed by the seals 5250, 5270, and the pressurizedfluid disposed in the space 5330 discourages or prevents leaked workingfluid in the vent 5290 from further leaking past the seals 5250, 5270.Because the pressure in the inlet into the compressor 5150 is lower thanthe pressure in the space 5330, leaked fluid will flow back to the inletrather than leaking past the hydraulic packing.

According to various embodiments, the seal 5200 may be modified byadding or removing various seals. For example, the compressor 5150includes one more seal between the compression chamber and the vent thanis included in the compressor 1000. In particular, in the compressor5150, four seals are disposed between the compression chamber 1020 andthe vent 5290 (i.e., the seals 1520, 1510, 5240, 5260), while theillustrated compressor 1000 has three such seals (i.e., seals 1520,1510, 1550). However, according to alternative embodiments greater orfewer such seals may be disposed between the compression chamber andvent without deviating from the scope of various embodiments. Forexample, one or more of the seals 1520, 1510, 5240, 5260 may be omitted.Alternatively, additional seals like the seals 5240, 5260 may extendbetween the collar 5220 and the face 5210 a of the casing 5210 tofurther reduce leakage from the compression chamber 1020, and the collar5220 and faces 5210 a,b may be radially expanded to provide space forsuch additional seals, preferably without axially elongating the overallmechanical seal. Additionally and/or alternatively, the seal 5200 may bemodified by adding a radial seal (e.g., like the seal 1550) between thecasing 5210 and shaft 1030 along the leakage path between the seals1510, 5240. Additionally and/or alternatively, the vent 5290 may bedisposed along the leakage path between different ones of the seals1520, 1510, 5240, 5260. For example, the vent may alternatively bedisposed in the leakage path between the inner face seal 5240 and theouter face seal 5260.

As shown in FIGS. 41 and 43, according to various embodiments, one ormore holes 1040 a extend axially through the entire rotor 1040 so as tofluidly connect opposite axial ends of the rotor 1040 radially inwardlyfrom the seals 1520. These holes 1040 a may prevent the rotor 1040 frombeing axially pushed against one axial end of the compression chamber1020 if compressed working fluid asymmetrically leaked past one of theseals 1520 on one axial end of the rotor 1040 to a greater extent thanat the opposite axial end of the rotor 1040. Additionally and/oralternatively, the fluid communication between the axial ends of therotor 1040 may be provided by extending a fluid passageway through theend plates 1545 of the casing 1010 (see FIG. 52), instead of through therotor 1040.

As shown in FIG. 52, according to various embodiments, a proximitysensor 1580 (e.g., contact or non-contact sensor, capacitive sensor,magnetic sensor, etc.) monitors the axial position of the rotor 1040relative to the end plates 1545 or other part of the casing 1010. Thesensor 1580 and associated controller (e.g., electronic control unit,analog or digital circuitry, a computer such as a PC, etc.) may causeone or more actions (e.g., an audio or visual alarm, deactivation of thecompressor) to occur if the sensed distance exceeds a predetermineddistance or falls below a predetermined distance

FIGS. 55-58 illustrate a compressor 2000 according to an alternativeembodiment. The compressor 2000 is generally similar to theabove-discussed compressors. Accordingly, a redundant description ofsimilar or identical components is omitted. The compressor 2000 includesa main casing 2010 that defines a compression chamber 2020, a driveshaft 2030, a rotor 2040 mounted to the drive shaft 2030 for rotationwith the drive shaft 2030 relative to the casing 2010, a gate 2050slidingly connected to the casing 2010 for reciprocating movement, and agate-positioning system 2060. The gate-positioning system 2060 of thecompressor 2000 differs from the gate-positioning systems of theabove-described compressors.

As shown in FIGS. 55-58, the gate-positioning system 2060 includes: agate-positioning-system casing 2070 mounted to the main casing 2010(e.g., via bolts or integral formation) (see FIGS. 56 and 58), a drivepulley 2080 mounted to the driveshaft 2030 for rotation with the driveshaft 2030, a cam shaft 2090 rotationally mounted to the casing 2070 forrelative rotation about a cam shaft axis that is parallel to an axis ofthe main drive shaft 2030, a driven pulley 2095 mounted to the cam shaft2090 for rotation with the cam shaft 2090 relative to the casings 2070,2010, a belt 2100 connected to the pulleys 2080, 2095, two cams 2110mounted to the camshaft 2090 for rotation with the camshaft 2090, camfollowers 2120 rotationally mounted to gate supports 2130 for rotationrelative to the supports 2130 about axes that are parallel to therotational axes of the shafts 2030, 2090, and springs 2140 that extendbetween the casing(s) 2070, 2010 and the gate supports 2130.

The gate supports 2130 mount to the gate 2050 to drive the reciprocatingmotion of the gate 2050. As shown in FIG. 57, the gate supports 2130pass through enlarged lower openings 2050 a in the gate 2050 and rigidlyattach (e.g., via a threaded connection, a retainer key or ring, aretainer pin 2135 (as shown in FIG. 57), etc.) to upper portions of thegate 2050 near an upper sealing edge 2050 b of the gate 2050. The loweropenings 2050 a are enlarged relative to the gate supports 2130 so thatthe gate supports 2130 do not contact lower portions of the gate 2050.According to various embodiments, extending the gate supports 2130through the enlarged lower openings 2050 a limits the effect thatthermal expansion/contraction has on the positioning of the seal 2050 bof the gate 2050 relative to the gate support 2130 position. Inparticular, thermal expansion of the gate 2050 below where the gate 2050mounts to the gate supports 2130 does not affect the positioning of thegate's seal 2050 b relative to the gate supports 2130. According tovarious embodiments, this provides more precise and accurate gate seal2050 b positioning relative to the rotor 2040 when the gate 2050thermally expands or contracts during use of the compressor 2000.

As shown in FIGS. 56 and 57, the gate supports 2130 slidingly mount tothe casing 2070 and/or 2010 via linear bearings 2137 (or other linearconnections such as bushings, etc.) to permit the gate supports 2130 tomove in the reciprocating direction of the gate 2050 (up/down as shownin FIGS. 56 and 57). An upper end of the springs 2140 abuts a springretainer portion of the casing 2070 and/or casing 2010. A lower end ofthe springs 2140 connects to the gate supports 2130 via spring retainers2150 or other suitable connectors. As a result, the compression springs2140 urge the gate supports 2130 and gate 2050 downwardly away from therotor 2040 and towards the cams 2110.

During operation of the compressor 2000, the drive shaft 2030rotationally drives the pulley 2080, which rotationally drives the belt2100, which rotationally drives the pulley 2095, which rotationallydrives the shaft 2090, which rotationally drives the cams 2110. Rotationof the cams 2110 drives the cam followers 2120, gate support 2130, andgate 2050 upwardly toward the rotor 2040 against the spring bias of thesprings 2140. The cams 2110 are shaped and the belt 2100 and pulleys2080, 2095 are timed so that the gate positioning system 2060 maintainsthe seal 2050 b of the gate 2050 proximate to (e.g., within 5, 4, 3, 2,1, 0.5, 0.3, 0.1, 0.05, 0.04, 0.03, 0.02, 0.01, 0.005, 0.004, 0.003,0.002, and/or 0.001 mm of) the rotor 2040 as the rotor 2040 rotatesduring operation of the compressor 2000. The gate-positioning system2060 therefore generally works in a similar manner as the gatepositioning system illustrated in FIG. 1, except that the relative rolesof the springs and cams are reversed in the compressor 2000 (i.e., thecams 2110 urge the gate 2050 toward the rotor 2040, rather than awayfrom it, and the springs 2140 urge the gate 2050 away from the rotor2040, rather than toward it).

In the gate-positioning system 2060 according to various non-limitingembodiments, a mass of the reciprocating components (e.g., the gate2050, gate supports 2130, cam followers 2120, portions of the springs2140 and retainers 2150) is kept relatively low to reduce the forcesneeded to drive such reciprocation. According to various embodiments,such reduction in reciprocating mass may facilitate higher compressor2000 operational speeds (in terms of RPMs) and/or smaller springs 2140and other structural components of the system 2060.

In the illustrated embodiment, the cam shaft 2090 is belt-driven via thepulleys 2080, 2095 and belt 2100. However, according to alternativeembodiments, the cam shaft 2090 may be driven by any other suitablemechanism for transferring rotation from the drive shaft 2030 to the camshaft 2090 (e.g., chain drive, gear drive, etc.) without deviating fromthe scope of various embodiments.

As shown in FIGS. 56-58, the casing 2070 encloses many of the componentsof the gate-positioning system 2060. In the illustrated embodiment, theonly working fluid leakage path to the ambient environment via the gate2050/casing 2010 interface is via the intersection between a hole 2070 ain the casing 2070 and the cam shaft 2090 on the side of the casing 2070where the cam shaft 2090 projects through the casing 2070 so that it maybe driven by the pulley 2095. As shown in FIG. 57, a hydraulic packing2170 seals this leakage path/intersection between the cam shaft 2090 andcasing 2070. According to various embodiments, the hydraulic packing2170 may be similar to or identical to the above-discussed hydraulicpacking 1590, and may comprise facing radial seals (e.g., similar to oridentical to the seals 1600, 1610) with a hydraulic fluid passage (e.g.,similar to or identical to the passage 1620) therebetween. The hydraulicpump 1380 may provide pressurized hydraulic fluid to the hydraulicpacking 2170 via a port/passageway (e.g., similar to or identical to theport/passageway 1630) that leads into the space between the seals. As aresult, the pressure within the hydraulic packing 2170 exceeds apressure within the casing 2070 so that fluids (e.g., working fluid thatleaked past the gate 2050 into the casing 2070 volume) do not leak outof or are discouraged from leaking out of the casing 2070. The casing2070 may be pressurized by working fluid that escaped from thecompression chamber 2020, and that pressure may prevent or discouragefurther leakage through that flow path.

Additionally and/or alternatively, as shown in FIG. 56, a vent passage2180 may fluidly connect the interior of the casing 2070 with the inlet(e.g., via the inlet manifold 2190 or a direct connection to the inletin the casing 2010). Such a vent passage 2180 may help to ensure that apressure in the casing 2070 remains below a hydraulic pressure in thehydraulic packing 2170 so as to further discourage working fluid in thecasing 2070 from leaking past the hydraulic packing 2170.

According to alternative embodiments, the hydraulic packing 2170 may bereplaced with any other suitable seal (e.g., conventional hermetic sealsthat are designed to seal rotating shafts where there is a significantpressure differential between opposing sides of the seal) or eliminatedaltogether (e.g., if the gate 2050's seal is sufficient) withoutdeviating from the scope of various embodiments.

According to an alternative embodiment, the casing 1010 and 2070 areaxially extended to entirely enclose the pulleys 2080, 2095 and camshaft 2090 such that only the main drive shaft 2030 of the compressor2000 extends from the casing 2010, 2070, requiring a single mechanicalseal like the seal 2170 between the drive shaft 2030 and elongatedcasing to hermetically seal the compressor 2000.

FIGS. 59-60 illustrate a compressor 3000 according to an alternativeembodiment. The compressor 3000 is generally similar to theabove-discussed compressor 2000. Accordingly, a redundant description ofsimilar or identical components is omitted. The compressor 3000 differsfrom the compressor 2000 by adding two additional sub-compressors thatare axially spaced from each other. Thus, the compressor 3000 comprisesthree sub-compressors 3000 a, 3000 b, 3000 c. The compressor 3000includes a main casing 3010 that defines three compression chambers 3020a, 3020 b, 3020 c, a drive shaft 3030, three rotors 3040 a, 3040 b, 3040c mounted to the drive shaft 3030 for rotation with the drive shaft 3030relative to the casing 3010, three gates 3050 a, 3050 b, 3050 cslidingly connected to the casing 3010 for reciprocating movement, and agate-positioning system 3060 that includes three cams 3110 a, 3110 b,3110 c mounted to the cam shaft 3090, three cam followers 3120 a, 3120b, 3120 c, three gate supports 3130 a, 3130 b, 3130 c, and three springs3140 a, 3140 b, 3140 c. The gate-positioning system 2060 of thecompressor 2000 differs from the gate-positioning systems of theabove-described compressors. Each of the respective sets of a, b, and ccomponents (e.g., compression chamber 3020 a, rotor 3040 a, gate 3050 a,cam 3110 a, cam follower 3120 a, gate support 3130 a, and spring 3140 a)work in substantially the same manner as the comparable components ofthe whole compressor 2000.

The inlet manifold 3500 of the compressor 3000 fluidly connects to theinlets of each sub-compressor 3000 a, 3000 b, 3000 c. According tovarious embodiments, the working fluid inlets of the threesub-compressors 3000 a, 3000 b, 3000 c fluidly connect to eachdownstream from the manifold 3500. Similarly, the compressed workingfluid outlets of the three sub-compressors 3000 a, 3000 b, 3000 c rejoinin the compressor's discharge manifold 3510. According to variousembodiments, check-valves are disposed in each sub-compressor'sdischarge outlets upstream from where the discharge passageways jointogether.

According to various embodiments, check-valves are also disposed in eachsub-compressor's inlet downstream from where the inlet flow pathdiverges toward respective sub-compressors 3000 a, 3000 b, 3000 c (e.g.,downstream or within the inlet manifold 3500) so as to discouragebackflow from one chamber 3020 a, 3020 b, 3020 c into another chamber3020 a, 3020 b, 3020 c during out-of-phase operation of thesub-compressors 3000 a, 3000 b, 3000 c.

As shown in FIGS. 59 and 60, the compression cycles of the compressors3000 a, 3000 b, 3000 c are 120 degrees out of phase with each other.Thus, when the sub-compressor 3000 a begins its compression cycle, thesub-compressor 3000 b is ⅓ of the way through its cycle, and thesub-compressor 3000 c is ⅔ of the way through its cycle. Positioning thesub-compressors 3000 a, 3000 b, 3000 c out of phase in this mannerreduces the maximum instantaneous torque that must be applied to thecompressor 3000, which may reduce the size/power/HP of the engine,motor, or other rotational driver being used to drive the drive shaft3030 of the compressor 3000. The 3-phase operation of the compressor3000 may also reduce vibrations as the reciprocating movement of thegate-positioning system are generally balanced across the threesub-compressors 3000 a, 3000 b, 3000 c. The 3-phase operation of thecompressor 3000 may also reduce pressure spikes downstream from thecompressor 3000 (e.g., in the discharge manifold 3510) because thecompressed fluid flow is divided into three sequential bursts for eachrevolution of the drive shaft 3030 (as opposed to a single larger burstin the compressor 2000). The 3-phase operation of the compressor 3000may also increase the strength of the casing 3010 and reduce therequired reinforcement of the casing 3010 around the gate because thesingle gate slot of the compressor 2000 is replaced with 3 gate slotswith reinforcing structure therebetween. The 3-phase operation of thecompressor 3000 may reduce the cost of the compressor 3000 because thenarrower gates 3050 a, 3050 b, 3050 c or rotors 3040 a, 3040 b, 3040 c(or other components of the compressor 3000) may be more easilyfabricated because they are not as long. The 3-phase operation of thecompressor 3000 may reduce the cost of the compressor 3000 becausebearings may be disposed between adjacent compression chambers 3020 a,3020 b, 3020 c, which can reduce drive shaft 3030 deflection, andfacilitate less expensive drive shafts 3030 and other components, whilestill maintaining tight tolerances between the rotor 3040 a, 3040 b,3040 c and casing 3010.

While the illustrated compressor 3000 includes three sub-compressors3000 a, 3000 b, 3000 c, the compressor may include greater or fewersub-compressors without deviating from the scope of various embodiments(e.g., n sub-compressors that operate out of phase by 360/n degrees fromeach other, where n is an integer greater than 1 and preferably lessthan 100 (e.g., 2, 3, 4, 5, 6, 7, 8, 9, 10)).

Alternatively, the multi-phase concept of the compressor 3000 may beimplemented using three discrete compressors (e.g., any of the abovediscussed compressors such as the compressors 1000, 2000, 5150) byconnecting their respective drive shafts (e.g., via direct co-axialmounting such that the compressors are axially spaced from each otheralong a common drive shaft, via gears, belts, etc.) such that thecompressors 1000, 2000, 5150 are out of phase from each other in thesame way that the above-discussed sub-compressors 3000 a, 3000 b, 3000 care out of phase with each other.

FIGS. 61-65 illustrate a compressor 4000 according to an alternativeembodiment. The compressor 4000 is generally similar to theabove-discussed compressor 2000, except that the compressor 4000 uses apivoting gate 4050, rather than a linearly reciprocating gate 1110.Accordingly, a redundant description of similar or identical componentsis omitted. The compressor 4000 includes a main casing 4010 that definesa compression chamber 4020 (see FIGS. 61-62), a drive shaft 4030rotationally mounted to the casing 4010, a rotor 4040 (see FIGS. 61-62)mounted to the drive shaft 4030 for rotation with the drive shaft 4030relative to the casing 4010, a gate 4050 mounted to a gate shaft 4052for common pivotal movement relative to the casing 4010 about a gateaxis 4055, a gate-positioning system 4060, a discharge manifold 4150 influid communication with an outlet 4160 into the compression chamber4020, and an inlet manifold 4170 in fluid communication with an inlet4180 of the compression chamber 4020.

As shown in FIGS. 61-62, the inlet 4180 passes through the gate 4050.This allows for a larger inlet 4180 area as well as a more efficient gasflow path. However, according to alternative embodiments, the inlet 4180may be spaced from the gate 4050 without deviating from the scope ofvarious embodiments.

As shown in FIGS. 63-65, the gate-positioning system 4060 includes a cam4110 mounted to the drive shaft 4030 for rotation with the driveshaft4030. An outer cam profile of the cam 4110 generally mimics a profile ofthe rotor 4040 (but may be modified to account forpivotal-position-based changes in the way the cam 4110 drives the camfollower 4120 relative to the gate 4050), a cam follower 4120 that abutsthe cam 4110 and is mounted to the gate shaft 4052 for common pivotalmovement with the shaft 4052 and gate 4050 relative to the casing 4010about the axis 4055 (see FIGS. 63-65), and a spring 4140 disposedbetween the casing 4010 and the gate 4050 to pivotally bias the gate4050 toward the rotor 4040. As the rotor 4040 rotates, thegate-positioning system 4060 keeps a seal edge 4050 a of the gateproximate to the rotor 4040. The spring 4140 urges the gate 4050 towardthe rotor 4040, while the cam 4110 and follower 4120 counter that forceso that the seal edge 4050 a closely follows the rotor 4040 surfaceduring operation of the compressor 4000.

The pivoting gate 4050 helps the gate 4050 to resist the pressure thatbuilds up on the compressed fluid outlet 4160 side of the gate 4050within the compression chamber 4020. As shown in FIGS. 61-62, theconvex, semi-cylindrical surface of the gate 4050 that is exposed tohigh pressures in the compression volume of the compression chamber 4020(the right side as shown in FIGS. 61 and 62) is concentric with the gateshaft 4052 and axis 4055. As a result, pressure loads are transferredthrough the gate 4050 directly to the shaft 4052 without urging the gate4050 to pivot. This direct force transfer through the shaft 4052 to thecasing 4010 may reduce gate 4050 deflection, and reduce the forcesneeded to reciprocally pivot the gate 4050 over each compression cycleof the compressor 4000, while keeping the seal edge 4050 a proximate tothe rotor 4040.

According to various embodiments, the gate 4050 and shaft 4052 may beintegrally formed.

In the illustrated embodiment, a torsion spring 4140 urges the gate 4050toward the rotor 4040. However, any other suitable force-impartingmechanism may alternatively be used without deviating from the scope ofthe present invention (e.g., a compression or tension spring mountedbetween the casing 4010 and a lever arm attached to the gate 4050 orshaft 4052 to impart torque on the shaft 4052 and gate 4050, a motor,magnets, etc.).

FIG. 66 illustrates a compressor 5000 according to an alternativeembodiment. The compressor 5000 is identical to the compressor 1000,except that the compressor 5000 uses a different type of gate supportguide 5075 than the gate support guide 1075 of the compressor 1000. Aredundant description of identical structures is omitted.

As shown in FIG. 66, the gate support guide 5075 is divided into threeparts, 5075 a, 5075 b, 5075 c. Guide parts 5075 a, 5075 c comprise gatesupport bushings or bearings 5080 that guide the gate supports 5050 topermit reciprocating linear motion of the supports 5050 (in the up/downdirection as illustrated in FIG. 66). The central guide part 5075 b ismounted to the casing 1010 (or integrally formed with the casing 1010).The central guide part 5075 b connects to the guide parts 5075 a, 5075 cvia linear bearings 5090. The linear bearings 5090 permit the outerguide parts 5075 a, 5075 c to move toward and away from the centralguide part 5075 b (i.e., along the arrows 5100 shown in FIG. 66, whichextend left/right as shown in FIG. 66). The linear bearings 5090 preventthe outer guide parts 5075 a, 5075 c from moving relative to the centralguide part 5075 b in a direction perpendicular to the arrows 5100 (i.e.,in a direction into/out of the page as shown in FIG. 66). The linearbearings 5090 are used to correct for relative thermal expansion ofdifferent parts of the compressor 5000 (e.g., between the gate supportguide 5075 and the gate support cross-arm 5055), which might otherwisecause the gate support bearings 5080 to push or pull the gate supports5050 in the direction of the arrows 5100 and cause the supports 5050 tobind against the bearings 5080.

According to various alternative embodiments, the linear bearings 5090are replaced with alternative linear movement devices that permit thegate supports 5050 to move in the direction of the arrows 5100. Forexample, thermal growth can be accounted for by slightly undersizing thegate support 5050 relative to the linear bearings 5080. Additionallyand/or alternatively, the linear bearings 5080 may be fitted intoslotted holes in the gate casing 5075 such that the linear bearings 5080can move axially (in the direction of the arrows 5100) if needed due tothermal growth while movement in a perpendicular direction (i.e., in thedirection into the page as shown in FIG. 66) is constrained oreliminated.

FIGS. 70-74 illustrate a compressor 6000 according to an alternativeembodiment. The compressor 6000 is similar to or identical to thecompressor 1000, except as explained below. A redundant description ofstructures and features of the compressor 6000 that are identical orsimilar to structures or features of the compressor 1000 is thereforeomitted.

As shown in FIGS. 70-73, the compressor 6000 adds a casing 6010 thatencloses many or all moving parts of the compressor 6000 other than thedrive shaft 6020 that extends outwardly from one or more ends of thecompressor 6000.

As shown in FIG. 73, an upper portion 6030 of the casing 6010 may beintegrally formed with the main casing that defines the compressionchamber 6040 of the compressor 6000. Inlet and discharge manifolds 6050,6060, respectively, may be integrally formed into the upper portion 6030of the casing 6010. The upper portion 6030 structurally supports thehydrostatic bearing 6070 and gate 6080, and may include reinforcingstructures to stiffen the casing and resist deflection caused bypressure from the bearing 6070 and gate 6080.

As shown in FIGS. 70 and 71, the casing 6010 also includes a lowerportion 6100 with an internal cavity that houses the springs 6110. Theupper portion 6030 may bolt or otherwise removably attach to the lowerportion 6100 so that the upper portion 6030 and main components of thecompressor 6000 may be removed from the lower portion 6100 (e.g., formaintenance or replacement). The springs 6110 may be removable as a unitalong with the upper portion 6030 and main components of the compressor6000. Alternatively, the springs may remain with the lower portion 6100when the upper portion 6030 is removed.

According to various embodiments, the lower portion 6100 may include asump for oil from the compressor's hydraulic and lubrication systemssuch that fluids reservoirs are provided within the casing 6010.

As shown in FIG. 70, the casing 6010 also includes cam covers 6130 thatenclose and protect the cams and cam followers (e.g., cams 1050 andfollowers 1060, as shown in FIG. 40). A lubrication distribution system6140 (e.g., an oil pump and oil-filled reservoir) connect via conduits6150 to the inside of the covers 6130 to apply (e.g., spray or drip)lubricant onto the cams and followers, and in particular the interfacebetween the cams and followers (shown in FIG. 39). In variousembodiments, this system may be configured to create an oil bath,wherein some portion of the cams and cam followers may be submerged inoil for part or all of their motion. The system may be configured tocreate an optimal oil level so as to maximize lubrication provided tothe cams and cam followers while minimizing negative effects such as oilsplashing, generation of bubbles within the oil, etc. While the system6140 is illustrated as being on the outside of the casing 6010 in FIG.70, the entire system 6140 and conduits 6150 may alternatively bedisposed inside the casing 6010. As shown in FIG. 72, rotational seals6160 seal the rotational interface between the shaft 6020 and covers6130. Such seals 6160 may comprise mechanical seals (e.g., rings). Theseals 6160 may comprise multi-part hydraulic seals like the seal 1500,6200 that provide a drain and hydrostatic over pressure to discourageworking fluid that may leak past the drive shaft into the inside of thecovers 6130 from leaking further into the ambient environment outsidethe covers 6130 and casing 6010.

As shown in FIG. 73, oil conduits 6170 in the upper portion 6030 mayfeed oil to the hydrostatic bearing 6070. The hydrostatic bearing 6070comprises to separate bearing pads 6070 a,b (shown on the right and leftin FIG. 73) that sandwich the gate 6070 therebetween (rather than asingle O or oval shaped bearing). The two-piece bearing 6070 mayfacilitate grinding of the bearing 6070 and gate 6080 to reduceclearances therebetween when the bearing 6070 and gate 6080 are insertedinto a matching slot in the upper portion 6030 of the casing 6010.

As shown in FIG. 74, a gate ring mechanical/hydraulic seal 6200surrounds the gate 6080 and seals an inside of the compression chamber6040 from the hydrostatic bearing 6070 and lower portion 6100 of thecasing 6010. The gate ring hydraulic seal 6200 operates in a similarmanner as the seal 1500 to isolate the compression chamber 6040 from anouter environment, except that the seal 6200 seals against thereciprocating gate 6080, rather than the rotating drive shaft. The seal6200 comprises, in sequential order from the compression chamber 6040toward the bearing 6070: a first seal 6210, a drain groove (e.g., avent) 6220, a second seal 6230, a hydraulic fluid groove 6240, and athird seal 6250. According to various embodiments, the seals 6210, 6230,6250 and grooves 6220, 6240 extend continuously around the entireperimeter of the gate 6080. The seals 6210, 6230, 6250 may each comprisesingle continuous seals such as O-rings, or may comprise multi-partseals that together form a complete perimeter around the gate 6080.

According to alternative embodiments, the seals 6210, 6230, 6250 andgrooves 6220, 6240 do not extend continuously around the gate 6080, butinstead are formed by two sets of seals and grooves, one set beingdisposed on the inlet side of the gate 6080 and one set being disposedon the outlet side of the gate 6080.

As shown in FIG. 74, the drain groove (e.g., vent) 6220 fluidly connectsto the inlet manifold 6050 via a fluid passageway 6280 so that workingfluid that leaks from the compression chamber 6040 past the first seal6210 is vented back into the low-pressure inlet manifold 6050 forreinjection back into the compression chamber 6040.

As shown in FIG. 74, the hydrostatic fluid groove 6240 is pressurized byhydraulic fluid (or other suitable fluid) that is pumped into the groove6240 via a fluid passageway 6290 from a source of pressurized fluid(e.g., hydraulic pump 1380).

As shown in FIG. 74, the seal 6200 includes a housing/body 6300 thatsupports the seals 6210, 6230, 6250 and grooves/vents 6220, 6240, anddefines portions of the passageways 6280, 6290. Other portions of thepassageways 6280, 6290 may be defined by the casing portion 6030 orother structures. The seal 6200 and its components are preferablyremovably inserted into place within the casing portion 6030 as a singleunit. As shown in FIG. 74, the seal 6200 is inserted into a mating slotin the casing portion 6030 from below. An additional seal ring 6310seals the interface between the body 6300 of the seal 6200 and thecasing 6030.

The operation of the seal 6200 is described with reference to FIG. 74.For the working fluid (e.g., natural gas being compressed) to leak outof the compression chamber 6040 via the opening through which the gate6080 extends, the fluid may leak between the seal 6210 and gate 6080. Ifthe working fluid leaks past the seal 6210, the fluid reaches the vent6220, which returns the fluid back to the low-pressure compressor inlet6050 via the passageway/port 6280, which is maintained at the pressureof the inlet 6050 via its fluid communication with the inlet 6050. Thearea between the second and third seals 6230, 6250 is pressurized byhydraulic fluid fed through the passageway 6290 and groove 6240 to apressure higher than the inlet 6050 pressure, which discourages orprevents the working fluid from further leaking past the seals 6230,6250 and groove 6240. Leaked working fluid leaks through the groove 6220and passageway 6280 back to the intake 6050, rather than past the seals6230, 6250 and groove 6240 because the inlet 6050 is at a significantlylower pressure than the groove 6240. Thus, leakage of the working fluidpast the seal 6200 is reduced or preferably eliminated.

According to various alternative embodiments, additional seals like theseals 6210, 6230, 6250 and corresponding vents like the vents 6220, 6240may be disposed along the leakage path between the first of such sealsand the last of such seals, which results in a plurality of drain vents6220 back to the inlet and/or a plurality of pressurized vents/grooves6240, with seals separating the different ones of the vents/grooves6220, 6240. According to various embodiments, the total number of suchseals along the leakage path may comprise from 3 to 50 seals.

According to alternative embodiments, the first seal 6210 and vent 6220may be eliminated so that the mechanical seal 6200 relies on thepressurized groove/vent 6240 to discourage leaks across the seal 6200.According to alternative embodiments, the third seal 6250 andvent/groove 6240 are eliminated, so that the mechanical seal 6200 relieson the vent 6220 to discourage further leakage past the seal 6230.

According to various embodiments, a flywheel may be added to one or bothends of the drive shaft 6020 to reduce torsional loads on the shaft 6020during operation of the compressor 6000.

According to various embodiments, any of the components or features(e.g., hydrostatic bearing 1300, mechanical seal 1500, compression ofmulti-phase fluids, etc.) of any of the above-described compressors(e.g., compressors 1000, 2000, 3000, 4000, 5000, 5150, 6000) may be usedin any of the other compressors described herein. For example, thedischarge manifold 1160 may be mounted to the outlet side 154 of thegate casing 150 of the compressor illustrated in FIG. 28 so as toreceive compressed fluid that is expelled through outlet ports 435.

The presently preferred embodiments could be modified to operate as anexpander. Further, although descriptions have been used to describe thetop and bottom and other directions, the orientation of the elements(e.g. the gate 600 at the bottom of the rotor casing 400) should not beinterpreted as limitations on embodiments of the present invention.

While various of the above-described embodiments comprise a rotarycompressor that relies on a rotor that is rigidly mounted to a driveshaft so that the rotor and drive shaft rotate together relative to thecompression chamber, various of the above-discussed features may be usedwith other types of compressors (e.g., rolling piston, screw compressor,scroll compressor, lobe, liquid ring, and rotary vane compressors)without deviating from the scope of these embodiments or the invention.For example, the above discussed hydrostatic bearing arrangement 1300can be incorporated into a variety of other types of compressors thatuse moving gates/vanes (e.g., rolling piston compressors, rotary vanecompressors, etc.) without deviating from the scope of such embodimentsor the invention.

While the foregoing written description of various embodiments of theinvention enables one of ordinary skill to make and use what isconsidered presently to be the best mode thereof, those of ordinaryskill will understand and appreciate the existence of variations,combinations, and equivalents of the specific embodiment, method, andexamples herein. The invention should therefore not be limited by theabove described embodiment, method, and examples, but by all embodimentsand methods within the scope and spirit of the invention.

It is therefore intended that the foregoing detailed description beregarded as illustrative rather than limiting, and that it be understoodthat it is the following claims, including all equivalents, that areintended to define the spirit and scope of this invention. To the extentthat “at least one” is used to highlight the possibility of a pluralityof elements that may satisfy a claim element, this should not beinterpreted as requiring “a” to mean singular only. “A” or “an” elementmay still be satisfied by a plurality of elements unless otherwisestated.

1-11. (canceled)
 12. A compressor comprising: a casing with an innerwall defining a compression chamber, an inlet leading into thecompression chamber, and an outlet leading out of the compressionchamber; a rotor coupled to the casing for rotation relative to thecasing; a gate movably coupled to one of the casing and rotor formovement relative to the one of the casing and rotor, the gatecomprising a sealing edge, the gate being operable to locate the sealingedge proximate to the other of the casing and rotor as the rotorrotates; and a hydrostatic bearing arrangement disposed between (1) thegate and (2) the one of the casing and rotor to reduce friction when thegate moves during operation of the compressor.
 13. The compressor ofclaim 12, wherein: the one of the casing and rotor comprises the casing;the gate is coupled to the casing for movement relative to the casing;the gate is operable to move relative to the casing to locate thesealing edge proximate to the rotor as the rotor rotates such that thegate separates an inlet volume and a compression volume in thecompression chamber; and the inlet and outlet are disposed on oppositesides of the sealing edge from each other.
 14. The compressor of claim13, wherein the hydrostatic bearing arrangement comprises: first andsecond inlet-side hydrostatic bearings disposed on an inlet side of thegate, the first and second inlet-side hydrostatic bearings being alignedand separated from each other along a direction of movement of the gate,and first and second outlet-side hydrostatic bearings disposed on anoutlet side of the gate, the first and second outlet-side hydrostaticbearings being aligned and separated from each other along a directionof movement of the gate.
 15. The compressor of claim 12, furthercomprising a drive shaft coupled to the casing for common rotation withthe rotor relative to the casing, wherein the rotor has a non-circularprofile. 16-20. (canceled)
 21. A compressor system comprising: aplurality of compressors, each compressor comprising: a casing with aninner wall defining a compression chamber, an inlet leading into thecompression chamber, and an outlet leading out of the compressionchamber, a rotor rotatably coupled to the casing for rotation relativeto the casing, and a gate coupled to the casing for movement relative tothe casing, the gate comprising a sealing edge, the gate being operableto move relative to the casing to locate the sealing edge proximate tothe rotor as the rotor rotates such that the gate separates an inletvolume and a compression volume in the compression chamber, the inletand outlet being disposed on opposite sides of the sealing edge fromeach other; and a mechanical linkage between the rotors of the pluralityof compressors, the mechanical linkage connecting between the rotorssuch that compression cycles of the plurality of compressors are out ofphase with each other.
 22. The compressor of claim 21, wherein theplurality of compressors comprises n compressors, and wherein themechanical linkage connects the rotors such that the compression cycleof each of the n compressors is out of phase with phase-wise adjacentones of the n compressors by 360/n degrees, and wherein 2≤n≤100.
 23. Thecompressor of claim 21, wherein the mechanical linkage comprises acommon drive shaft that extends through each of the plurality ofcompressors and is coupled to the rotors of each of the plurality ofcompressors for common rotation relative to the casings of each of theplurality of compressors.
 24. A compressor comprising: a casing with aninner wall defining a compression chamber, an inlet leading into thecompression chamber, and an outlet leading out of the compressionchamber; a drive shaft and rotor rotatably coupled to the casing forcommon rotation relative to the casing such that when the rotor isrotated, the compressor compresses working fluid that enters thecompression chamber from the inlet, and forces compressed working fluidout of the compression chamber through the outlet; and a mechanical seallocated at an interface between the drive shaft and casing where thedrive shaft passes through the casing, the mechanical seal comprising:first, second, and third seals disposed sequentially along a leakagepath between the drive shaft and casing rotor, a source of pressurizedhydraulic fluid, and a hydraulic fluid passageway that connects thesource to a space along the leakage path between the second and thirdseals so as to keep the space pressurized with hydraulic fluid.
 25. Thecompressor of claim 24, wherein the mechanical seal further comprises avent disposed between the first and second seals, the vent being fluidlyconnected to the inlet so as to direct working fluid that leaks from thecompression chamber past the first seal back to the inlet.
 26. Thecompressor of claim 25, wherein the mechanical seal is configured tomaintain a hydraulic pressure between the second and third seals that ishigher than a pressure at the inlet.
 27. The compressor of claim 24,wherein the first, second, and third seals each comprise radial shaftseals extending radially between the drive shaft and the casing, whereineach of the shaft seals abuts an outer circumferential surface of thedrive shaft.
 28. The compressor of claim 24, wherein the first, second,and third seals each comprise face seals extending axially between thecasing and an axial face that rotates with the drive shaft. 29-32.(canceled)
 33. A compressor comprising: a casing with an inner walldefining a compression chamber, an inlet leading into the compressionchamber, and an outlet leading out of the compression chamber; a rotorrotatably coupled to the casing for rotation relative to the casing suchthat when the rotor is rotated, the compressor compresses working fluidthat enters the compression chamber from the inlet, and forcescompressed working fluid out of the compression chamber through theoutlet; a gate coupled to the casing for reciprocating movement relativeto the casing, the gate comprising a sealing edge, the gate beingoperable to move relative to the casing to locate the sealing edgeproximate to the rotor as the rotor rotates such that the gate separatesan inlet volume and a compression volume in the compression chamber; anda mechanical seal located at an interface between the gate and casing,the mechanical seal comprising: first, second, and third seals disposedsequentially along a leakage path between the gate and casing, a sourceof pressurized hydraulic fluid, and a hydraulic fluid passageway thatconnects the source to a space along the leakage path between the secondand third seals so as to keep the space pressurized with hydraulicfluid.
 34. The compressor of claim 33, wherein the mechanical sealfurther comprises a vent disposed between the first and second seals,the vent being fluidly connected to the inlet so as to direct workingfluid that leaks from the compression chamber past the first seal backto the inlet.
 35. The compressor of claim 33, wherein the first, second,and third seals are all supported by a removable housing, such that thefirst, second, and third seals and housing can be installed into thecasing as a single unit.
 36. The compressor of claim 33, wherein themechanical seal comprises n sequential seals along the leakage pathbetween the gate and casing, wherein 3≤n≤50, wherein n includes thefirst, second, and third seals, wherein one or more spaces betweenadjacent ones of the seals are filled with pressurized hydraulic fluid,and wherein one or more spaces between adjacent ones of the sealscomprise a vent that is fluidly connected on the inlet.